WO2002081094A1

WO2002081094A1 - A gearbox for a centrifuge, such as a decanter centrifuge

Info

Publication number: WO2002081094A1
Application number: PCT/DK2002/000224
Authority: WO
Inventors: Jan Michelsen; Raymond John Hicks
Original assignee: Centriquip Limited
Priority date: 2001-04-04
Filing date: 2002-04-04
Publication date: 2002-10-17
Also published as: WO2002081094B1

Abstract

A gearbox for a centrifuge, such as a decanter centrifuge with a rotatable bowl (1) connected to a housing (31) of the gearbox, is driven at a first speed by a first external driving means (11) and a helical conveyer (2) coaxially arranged for rotating therein at a second rotational speed. Said gearbox is comprising a three-stage epicycloidal gear train. The relative speed difference between the bowl and the conveyer is by means of the gearbox according to the invention achieved with smaller internal and external losses than hitherto known. Just by reversing the direction of the conveyer sun wheel the bowl can be brought to run faster or slower than the conveyer.

Description

A gearbox for a centrifuge, such as a decanter centrifuge

Field of the invention:

The present invention relates to a gearbox for a centrifuge, such as a decanter centrifuge with a rotatable bowl (1) connected to a housing (31) of the gearbox driven at a first speed by a first external driving means (11) and a helical conveyer (2) coaxially arranged for rotating therein at a second rotational speed.

Background of the invention:

A decanter centrifuge generally includes a rotatable bowl having a coaxially mounted screw conveyer therein. During operation the bowl is rotated at a constant but variable speed in order to create a centrifugal force to separate a fluid feed mixture into its constituent parts.

The heavier portion of the feed, typically called solids because of its, at least partially, conveyable nature, collects on the inner surface of the bowl due to centrifugal force. The screw conveyer is rotated at a relative speed with respect to the bowl. This differential rotation creates a differential action between the flights of the screw and the bowl wall resulting in the conveyance of the solids along the bowl wall.

This differential speed can be varied during the operation of the centrifuge depending on certain parameters and the desired output qualities of the separated constituent parts of the feed mixture. The light or liquid part of the feed moves radially inward of the heavier solids as a result of the centrifugal force. Thereafter, the separated heavy and light matters are separately discharged, typically from opposite ends of the bowl.

For very stable processes, the conveying speed can be kept constant. Such centrifuges have a simple conveyer drive comprising a cycloid (i.e. planetary) type of gear transmission between the conveyer and the bowl, with the reaction member (i.e. sun wheel shaft) fixed to the machine frame. In case of torque overload, a maximum load mechanism releases the fixture and lets the reaction member rotate with the bowl, thus reducing the conveying speed to zero, after which the centrifuge can be cleaned and restarted after re- engagement of the release mechanism. In most cases, the transmission comprises at least two gear stages, because one stage is not capable of dealing with the small ratio of movement between the bowl and the conveyer at the same time as having a large ratio of movement between the bowl and the stationary reaction member, (sun wheel shaft) .

For a traditional two-stage planetary gearbox with fixed or braked sun wheel, the conveyer is turned the same way as the sun wheel relative to the bowl, i.e. the conveyer runs at a slower speed than the bowl. The acceleration of the feed by the surface of the liquid occurs therefore in a zone extending from the conveyer feed ports towards the cylindrical end of the bowl. In some applications, this is a disadvantage to the separation process.

One recent prior art is shown in WO 91/10846, which provides a solution to this disadvantage, by one of the gear stages having two sets of meshed planet wheels supported by the same carrier. In this way, the conveyer is turned in the opposite direction to the bowl, i.e. the conveyer will run at a higher speed than the bowl, and the acceleration zone will be directed towards the conical end of the bowl.

As an improvement to the fixed reaction member concept above, the reaction member may be engaged by a fixed ratio pulley transmission, thus giving a means of changing the conveying speed by changing the pulleys. This adjustment can only be done, when the centrifuge is at a standstill.

A disadvantage of the above mentioned fixed-ratio solutions is that it cannot be changed while the centrifuge is operating, and that, if the speed of the bowl is decreased (in order to ease the G-force on the solids and thereby ease the conveying- out of the solids) , the speed of the conveyer decreases at the same rate as the bowl speed, thus reducing the speed of removal of solids from the bowl. It is desirable to have the conveying speed constant, or to increase, during such relief- actions, and applying a brake mechanism to the sun wheel provides a solution to this.

The brake mechanism causes the reaction member (sun wheel) to rotate at a speed between zero and the speed of the bowl, thereby adjusting the conveying speed between maximum speed

(given by the transmission ratio of the gearbox) and zero. The braking, however, causes power to leave the rotor system, and it will dissipate from the brake as heat. If a conveying speed close to zero is needed, the speed of the reaction sun wheel will be high (close to the bowl speed) , and the resulting power loss in the brake will be similarly high.

Over the years, a wide variety of variable conveying speed drives have been developed, including types depending on the gearbox transmission concept (i.e. variable hydraulic or electric (motor or eddy-current) braking arrangements attached to the reaction member) .

One such example is described in EP 0 493 009 Al, which covers a recent improvement of the conventional concept. In this attempt, the power is not lost by braking, but is recycled electronically back to the grid. A major disadvantage of this solution is that the wave shape of the recycled power is imperfect, causing noise in the power grid.

A common disadvantage of the above named types of systems is that the braking action causes a power loss that is either fully lost (by braking and dissipation as heat) or alternatively can be partly recovered only by means of complicated methods.

Danish patent application No. 5033/84 describes a 2-train differential gear construction having the ability to drive the conveyer without the speed being dependent of the bowl speed. One gear train of this construction is used as a reference to the rotating housing connected to the bowl, while the other train is referencing the conveyer. In this way, by fixing one of the sun wheels, a movement of the other sun wheel will turn the conveyer relative to the bowl independent of the rotor speed. In effect, this means that the whole ratio between the conveyer sun wheel and the movement of the conveyer relative to the bowl is made in one stage. Such a stage will have to be made with gears of considerable width to handle the full torque, and at the same time, the speed of the sliding tooth surfaces is high because of the required ratio of the transmission. High speed and high torque at the same time means high internal losses, and that a large amount of power is cycled between the two trains within the gearbox.

The application of this prior art was abandoned because of the high power losses that were revealed through the detailed design calculations, which made it commercially unattractive.

However, new detailed calculations have revealed that the internal losses can be reduced considerably by introducing a third, high-torque gear train between the conveyer and the driving member of the differential train. In this way, the simultaneous high-torque/high-speed conditions can be avoided, and the internal losses become comparable to the losses known from the traditional 2-stage gear applications.

A major advantage of this design is that the conveying power is provided directly from the input sun wheel to provide a movement of the conveyer relative to the bowl, without any braking taking place. The exterior losses have therefore been eliminated, and furthermore the variable control of the conveying speed can be done by standard equipment, i.e. a standard electrical motor controlled by a programmable inverter drive .

Object of the invention.

The object of the invention is to provide a gearbox for a centrifuge of the kind referred to in the opening paragraph by which it is possible to establish a relative speed difference between the bowl and the conveyer with smaller losses to the environment than hitherto known.

A second object of the invention is to provide such a gearbox by which it is possible to establish a relative speed difference between the bowl and the conveyer without using a braking action.

A third object of the invention is to provide such a gearbox by which it is possible to establish a relative speed difference between the bowl and the conveyer with smaller internal losses than previously known. Summary of the invention.

The present invention provides an efficient and uncomplicated way of operating a decanter centrifuge, firstly, because the efficiency of the gearbox of the invention is higher than the conventionally applied transmissions, and secondly because the power supplied to the conveyer is positive (no braking) . The conveyer speed control circuit (if any) is straightforward and independent of the bowl speed, and a direct reaction member torque reading is available for independent, static measurement of the torque. It comprises the most power- effective conveyer drive system of today.

The present invention is comprising a three stage epicyclic gear train, wherein two of the stages are combined to act as a double differential transmission input step, the input shaft of which is connected to a second external driving means, the reference shaft of which being connected to an external reaction member, and the output shaft of which being connected to the input shaft of the third gearbox stage, the output reference member of the differential stage being connected to the housing.

The built-in oil pump is based on the centripetal pump principle, and is therefore dependent on the direction of rotation to build up lubrication pressure. As the operation for certain applications for a centrifuge requires a change of direction of rotation, the pump is designed to act in both directions of rotation.

Because of the number of gear trains in the gearbox configuration, the layout constitutes a 4-member mechanism, which are: the input shaft, the output shaft connected to the conveyer, the rotating reaction member connected to the bowl, and the stationary reaction member connected to the machine frame . Because of this configuration it is possible to directly measure the reaction torque on the stationary reaction member by means of a simple sensor, thereby creating a directly torque-proportional control signal to the control circuit. In this way, the torque signal will be undisturbed by the electrical couplings in the inverter circuit, that occur in a conventional control system.

Brief description of the drawings.

The invention is illustrated by the following 5 drawings showing some of the various arrangements that are feasible :

Fig. 1 shows a general arrangement of a decanter centrifuge with a differential gear conveyer drive according to the invention .

Fig. 2 shows a first embodiment of the invention, a cross- sectional view of an alternative 3-train compound/differential gear transmission and oil pump.

Fig. 3 shows a second embodiment of the invention, a cross- sectional view of a 3-train planetary/differential gear transmission and oil pump.

Fig. 4 shows details of a lubrication pump according to the invention.

It must be emphasised that the above drawings are describing some, but not all of the possible designs for a differentially driven high torque conveying centrifuge gearbox. Alternative arrangements based on differential gear technology may give similar advantages. Detailed description of the drawings.

Fig. 1 shows a general arrangement of a decanter centrifuge with differential gear conveyer drive, comprising a bowl rotatably supported by bearings 12, and a helical conveyer 2 mounted therein, supported by bearings 13. At one end a shaft 23 extending from the conveyer 2 is connected to the bowl 1 through a differential gear transmission 7. An input member 15 of said gear transmission 7 is connected via a pulley transmission 20 to a variable speed motor 10, and a stationary reaction member of a gear transmission 14 is connected to a torque measurement device 8. A motor 11 drives the bowl 1 via a pulley transmission 9 providing a fixed speed of the bowl 1.

By this arrangement the conveyer 2 is brought to rotate within the bowl 1 at a speed slightly different than the bowl, thus creating a screw conveying action towards the conical end 5 of the bowl 1 to any matter 24 deposited onto the inner wall of the bowl 1. The conveying action requires quite a high driving torque, which is provided by the differential gear transmission 7.

The decanter centrifuge functions in the following way:

The slurry or suspension 21 that is to be separated is fed to the centrifuge through a feed pipe 3 and a feed chamber 4 in the interior of the conveyer. By centrifugal force, the feed is contained within the bowl 1 forming an annular liquid ring 19, from which the solids particles are precipitated towards the inner face of the bowl 1 forming a cake 24, which is subsequently conveyed towards the conical end 5 of the bowl 1 by the conveyer 2. Here, the cake compresses and dewaters during passage of the dry part of the bowl and leaves the centrifuge through apertures 16 in the bowl wall. At the other end of the bowl 1 the cleaned liquid 6 leaves the bowl 1 through another set of apertures 17. As the cake content in the bowl 1 increases, the torque on the differential gear transmission reaction member and drive motor shaft 10 increases, causing an inverter 18 to control the conveyer speed relative to the bowl in such a way as to stabilise the cake quality.

Fig. 2 shows a cross-sectional view of a differential gear conveyer drive, comprising a gearbox housing 31 attached to the bowl of the centrifuge 1 and an output shaft 23 attached to the conveyer of the centrifuge.

The transmission gearing takes the form of a 3-stage epicycloidal gear train.

The output stage 25 comprises a number of compound planet gears 29 (with a different number of teeth on the two gear meshes 39 and 40) mounted by bearings on a carrier 26, and the mesh 40 meshing with a ring gear 32 fast to the output shaft 23, and the mesh 39 meshing with a ring gear fast to the gear housing 31. The two ring gears having same number of teeth, one revolution of the carrier causes the output shaft 23 to move relative to the housing an amount given by the difference of teeth numbers on the compound planet gear meshes 39 and 40.

The input comprises a paired differential planetary gear transmission 33, shown as two stages of planetary gears of similar transmission ratio.

The input shaft 15 is connected to input gear mesh 37 meshed with a number of planet gears 41 mounted on a carrier 43. The planet gears mesh on the other side with a ring gear 36 made fast to the compound carrier 26.

The differential reference shaft 14 is connected to gear mesh 44 meshed with a number of planets 42 similar to the planets

41. The planet gears 42 mesh with a ring gear fast to the gear housing 31 .

On the differential reference shaft 14, which is held stationary during operation, a centripetal oil pump 38 extends towards the inner wall of the housing 31 to a distance allowing the scooping point 45 to scoop the oil into the channel 46 and through oil muffs (not shown) into the distribution holes 47 leading the oil to various spray positions 48.

The stationary reaction member is connected to a torque measurement device 8 (Fig. 1) to monitor the torque of the conveyer-bowl frictional interaction.

An inverter control system 18 (Fig. 1) is attached to the conveyer drive motor 10 in order to provide adjustment of torque level and relative speed in accordance with a pre-set ramp controller.

To explain the function of the gearbox, first imagine that the housing 31 and the reaction shaft 14 are kept stationary. If both meshes of the planet gears 42 are stationary, the planet wheel itself will be stationary, and from that, the carrier 43 will also be stationary.

Turning of the input shaft 15, therefore, will cause the ring gear 36 to turn the opposite way at a smaller rate given by the geometry, thus turning the carrier 26 of the compound stage. As the compound planet mesh 39 is meshing with a ring gear fast to the (stationary) housing 31, the movement will cause the output shaft 23 to turn relative to the housing and thereby to the bowl 1. The rotation of the conveyer is therefore directly proportional to the rotation of the input shaft.

Next, imagine that the housing 31 is rotating at high speed, while the reaction shafts 14 as well as the input shaft 15 are kept stationary. The planet gear 42 will then be rotated by the meshing ring gears fast to the housing 31, thus rotating the carrier 43 at a lower rate than the housing. The planet gear 41 will take on the same speed as 42, because they are mounted on a common carrier. The ring gear 36 and the carrier 26 fast with it will then move at the same speed as the housing 31, and with no relative movement between housing 31 and carrier 26 the output shaft 23 will have same speed as the housing 31. Thus zero speed on the input shaft 15 results in zero relative speed between housing 31 attached to bowl 1 and output shaft 23 attached to the conveyer.

A particular detail that is important to the reduction of the internal losses in the gear transmission is the oil pump 38. In a conventional 2-stage epicycloidal gearbox, the gearbox is filled to a level (when the gearbox rotates) inside of the planet centres, in order to provide lubricant to the highly loaded planet (journal) bearings. The oil is therefore subject to centrifugal forces, creating a high pressure near the wall of the gearbox. When a pair of teeth in a ring gear and a planet gear engages, the oil has to be pumped to the sides (distance: half a gear width) against this pressure, thereby causing a considerable loss that is dissipated as heat.

In the present design, the lubrication is provided by a centripetal pump 38, which takes advantage of the high rotating speed difference between the stationary reaction member and the rotating gearbox housing containing the lubricant by scooping the oil into a channel according to a well established technology. The centripetal pump delivers the pressurised oil via an oil transfer bushing/bearing to bores through the central members (shafts) , from which it is directed to different supply points for lubrication of journal as well as roller- and ball bearings, and teeth (by spray) . In this way, the oil level can be kept outside the pitch diameter of the ring gears, thus eliminating the considerable power loss caused by the gear teeth displacing the oil from the gear mesh.

Fig. 3 shows a cross-sectional view of a differential gear conveyer drive, similar to the arrangement of fig. 2, the difference being that the high-torque stage 25 is replaced by a planetary gear stage 55. The reason for this is that the losses caused by the gear mesh in the output stage can be lowered a little further, at the expense of a higher number of parts, higher weight of the gearbox, and higher bearing loads.

The output stage 55 comprises a number of planet gears 59 mounted by bearings on a carrier 50 fast with the output shaft 23, meshing on the one hand with a ring gear 52 fast to the housing 31, and on the other hand with a sun gear 51 fast to the differential input ring gear 56.

Apart from this, the arrangement is similar to fig. 2, and the function is easily seen to be similar as well.

Fig. 4 shows a centripetal oil pump with the ability to provide lubricant regardless of the direction of rotation of the centrifuge. The pump 38 is provided with two (sets of) pumping channels, one (62) having the aperture oriented for clockwise rotation of the bowl, and the other (63) oriented for anticlockwise rotation. The channels combine into a common channel 46, and the intersection is equipped with a flap-valve 65, which acts as a check-valve in such a way that only the pressurised channel is open and able to supply lubricant to the gearbox. (In the shown position the bowl rotates anticlockwise) . Description of the preferred embodiment.

Fig. 2 describes the preferred embodiment. This design has been developed to have short length, because in this particular case, the gear transmission is mounted inside of the main bearings of the centrifuge.

It is clear from the theory of rotor vibrations, that addition of weight and distance between the bearings lower the critical frequency of a given rotor, thus reducing the maximum operational speed of the rotor. As the efficiency of a centrifuge is proportional to the square of the operational speed, it will be clear that a reduction would result in a low-efficiency centrifuge.

The drive can be arranged in a number of configurations, depending on the degree of control needed, naturally as a cost reducing option if full control is not needed.

As a first simplification, the torque measurement device on the static reaction member can be omitted, and the torque measurement can be (indirectly) measured in the inverter circuit. This means that the measurement will be less accurate and probably delayed in relation to the direct measurement.

Just omitting the torque control system as a whole, just connecting a drive motor to the input sun wheel shaft and running the motor at a fixed speed can make a further simplification. To protect the gear mechanism against torque overload, a torque-release coupling can be connected to the stationary reaction member and should release the member in case of torque overload. Setting a max-current limit in the motor drive system may also do it. (In both cases a signal should be generated to serve as a switch for stopping the feed pump) . The same action as above can be obtained by driving the input shaft via a pulley transmission to the main drive motor. In this case the main drive motor has to be sized to deliver the conveyer drive power as well as the main drive power, and as an option the pulley drive could be of a variable ratio type to obtain variable speed on the conveyer drive. This can further be made automatically variably controlled by using a torque measurement signal from the stationary reaction member as control parameter for the transmission ratio, but this would be a quite complicated, bulky and expensive solution.

It must be emphasised that the above descriptions do not prevent the gearbox from being mounted in alternative arrangements relative to the centrifuge. By redesign of the attachment members it is possible to adapt the design for mounting in the other end of the centrifuge, inside the main bearings, and in both ends, outside of the main bearings. Further, by change of the gear trains to adapt wider gears in combination with smaller diameters of the components, it is possible to adapt the design for mounting into the conveyer body, thus minimising the use of space in the centrifuge.

The advantages regarding energy efficiency can be illustrated by the following Tables 1, 2 and 3 extracted from the dimensioning calculations of a gearbox according to the invention compared to a conventional gearbox.

Both gearboxes are mounted on centrifuges doing 4200 rpm, and they are compared at a small and a large relative speed and maximum torque capacity. They have comparatively equal overall transmission ratios (60 and 43.6)

Gear meshing losses are considered to be of same magnitude for both solutions, but the oil displacement loss from being pumped are only included on the conventional design. Table 1. Max conveying power transferred at max. torque

Table 2. Power loss

Table 3. Power loss in relation to transferred power (efficiency)

As described above, a further power loss is generated by both systems from the way the conveying power is handled.

In the system according to the invention, power is always added to the sun wheel by a motor controlled by an inverter, and the only loss is therefore the efficiency loss of the motor and control system, which is comparatively low in modern systems (i.e. less than 15%).

In a conventional variable system with regeneration, the power taken out by braking the sun wheel is (except for mechanical losses) returned to the grid, and the regeneration itself (i.e. by applying a motor as generator and by adjustment of frequency in a converter) is generating a loss from the current in the motor windings (i.e. 25%) . In a conventional variable system without regeneration, all the power taken out by braking the sun wheel is lost, as there is no system available for regeneration (i.e. adjustment of frequency) and returning it to the grid, and the braking itself (i.e. by applying a motor as generator) is generating a further loss from the current in the motor windings (i.e. 15%) .

In table 4 below, the added losses from the power handling are exemplified.

Table 4. Conveyer power handling loss from motor and control system.

It will be clear from the above that high torque in combination with high bowl speeds require solutions different from the conventional, particularly at low differential speeds, as with most applications with variable conveying speed, and the present invention has all the properties to improve this application considerably.

The present invention has both the ability of running faster as well as slower conveyer, just by reversing the direction of movement of the conveyer drive sun wheel .

Claims

What is claimed is:

1. A gearbox for a centrifuge, such as a decanter centrifuge with a rotatable bowl (1) connected to a housing (31) of the gearbox driven at a first speed by a first external driving means (11) and a helical conveyer (2) coaxially arranged for rotating therein at a second rotational speed, wherein the gearbox is comprising a three stage epicyclic gear train, wherein two of the stages are combined to act as a double differential transmission (33) input step, the input shaft of which (15) is connected to a second external driving means (10) , the reference shaft of which (14) being connected to an external reaction member, and the output shaft of which (36; 56) being connected to the input shaft (26) of the third gearbox stage (25;55), the output reference member of the differential stage (42) being connected to the housing (31) .

2. A gearbox according to claim 1, wherein the input shaft (15) is connected to a first sun gear (37) meshing with a first number of planet gears (41) mounted on a first carrier (43) upon which also is mounted a second number of planet gears (42) meshing with a first ring gear (28a) fast to the housing (31) and with a second sun gear (44) on a differential reference shaft (14) .

3. A gearbox according to claim 1 or 2 , wherein the gearbox is having an output stage comprising at least one planetary gear transmission (25) connected to the conveyer (2) and having a third number of planet gears (39; 40) mounted on a second joint carrier (26) with a ring gear (36) meshing with the first number of planet gears (41) mounted on the first carrier (43).

4. A gearbox according to claim 3, wherein the third number of planet gears consists of at least one first planet gear (39) having a first number of teeth and at least one second planet gear (40) , which has a different second number of teeth and is mounted on a joint axle on the second joint carrier (26), the first planet gear (39) meshing with a ring gear (32) fast to an output shaft (23) connected to the conveyer (2), and the second planet gear meshing with a ring gear (28b) fast to the housing (31) .

5. A gearbox according to claim 4, wherein the first numbers of teeth on the first planet gear (39) is lesser than the second number of teeth on the second planet gear (40) .

6. A gearbox according to claim 1 or 2 , wherein the gearbox is having an output stage comprising a planetary gear transmission (55) having a third number of planet gears (59) mounted on a second carrier (50) fast with an output shaft (23) connected to the conveyer (2), the third number of planet gears (59) meshing with a ring gear (52) fast to the housing (31) and with a sun gear (51) fast to a ring gear (56) meshing with the first number of planet gears (41) mounted on the first carrier (43) .

7. A gearbox according to each of the claims 1 - 6, wherein a centripetal pump (38) is arranged in relation to a stationary part of the decanter centrifuge inside the gear housing (319, the housing providing holding space for the lubricant in such a way as to keep the gear mesh free of the relative oil level inside the housing when the housing is rotating at its operating speed.

8. A gearbox according to claim 7, wherein the centripetal pump is able to pump oil regardless of the direction of rotation of the housing.

9. A gearbox according to each of the claims 1 - 8, wherein the differential reference shaft (14) is driven by a constant speed electric motor, said motor having at least one set of poles .

10. A gearbox according to each of the claims 1 - 8, wherein the differential reference shaft (14) is driven by an inverter controlled electric motor, said control consisting of a torque feedback circuit to enable constant filling with sediment in the centrifuge.

11. A gearbox according to each of the claims 1 - 8, wherein the differential reference shaft (14) is driven by an inverter controlled electric motor, said torque feedback being provided by a torque arm measurement on the stationary reaction member.