US20130052072A1 - Lubrication of screw machines - Google Patents

Lubrication of screw machines Download PDF

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Publication number
US20130052072A1
US20130052072A1 US13/578,348 US201113578348A US2013052072A1 US 20130052072 A1 US20130052072 A1 US 20130052072A1 US 201113578348 A US201113578348 A US 201113578348A US 2013052072 A1 US2013052072 A1 US 2013052072A1
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Prior art keywords
rotors
expander
working fluid
rotor
wet steam
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English (en)
Inventor
Ian Kenneth Smith
Nikola Rudi Stosic
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City University of London
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City University of London
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Assigned to THE CITY UNIVERSITY reassignment THE CITY UNIVERSITY ASSIGNMENT OF ASSIGNORS INTEREST (SEE DOCUMENT FOR DETAILS). Assignors: SMITH, IAN KENNETH, STOSIC, NIKOLA RUDI
Publication of US20130052072A1 publication Critical patent/US20130052072A1/en
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    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01C—ROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
    • F01C21/00—Component parts, details or accessories not provided for in groups F01C1/00 - F01C20/00
    • F01C21/04—Lubrication
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01C—ROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
    • F01C1/00—Rotary-piston machines or engines
    • F01C1/08—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing
    • F01C1/12—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of other than internal-axis type
    • F01C1/14—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons
    • F01C1/16—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing of other than internal-axis type with toothed rotary pistons with helical teeth, e.g. chevron-shaped, screw type
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01C—ROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
    • F01C17/00—Arrangements for drive of co-operating members, e.g. for rotary piston and casing
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01C—ROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
    • F01C21/00—Component parts, details or accessories not provided for in groups F01C1/00 - F01C20/00
    • F01C21/04—Lubrication
    • F01C21/045—Control systems for the circulation of the lubricant
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01K—STEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K25/00—Plants or engines characterised by use of special working fluids, not otherwise provided for; Plants operating in closed cycles and not otherwise provided for
    • F01K25/005—Plants or engines characterised by use of special working fluids, not otherwise provided for; Plants operating in closed cycles and not otherwise provided for the working fluid being steam, created by combustion of hydrogen with oxygen
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01K—STEAM ENGINE PLANTS; STEAM ACCUMULATORS; ENGINE PLANTS NOT OTHERWISE PROVIDED FOR; ENGINES USING SPECIAL WORKING FLUIDS OR CYCLES
    • F01K25/00—Plants or engines characterised by use of special working fluids, not otherwise provided for; Plants operating in closed cycles and not otherwise provided for
    • F01K25/04—Plants or engines characterised by use of special working fluids, not otherwise provided for; Plants operating in closed cycles and not otherwise provided for the fluid being in different phases, e.g. foamed
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C15/00—Component parts, details or accessories of machines, pumps or pumping installations, not provided for in groups F04C2/00 - F04C14/00
    • F04C15/0088—Lubrication
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2/00—Rotary-piston machines or pumps
    • F04C2/08—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C29/00—Component parts, details or accessories of pumps or pumping installations, not provided for in groups F04C18/00 - F04C28/00
    • F04C29/04—Heating; Cooling; Heat insulation
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16N—LUBRICATING
    • F16N15/00—Lubrication with substances other than oil or grease; Lubrication characterised by the use of particular lubricants in particular apparatus or conditions
    • F16N15/04—Lubrication with substances other than oil or grease; Lubrication characterised by the use of particular lubricants in particular apparatus or conditions with water
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16N—LUBRICATING
    • F16N7/00—Arrangements for supplying oil or unspecified lubricant from a stationary reservoir or the equivalent in or on the machine or member to be lubricated
    • F16N7/36—Arrangements for supplying oil or unspecified lubricant from a stationary reservoir or the equivalent in or on the machine or member to be lubricated with feed by pumping action of the member to be lubricated or of a shaft of the machine; Centrifugal lubrication
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F01—MACHINES OR ENGINES IN GENERAL; ENGINE PLANTS IN GENERAL; STEAM ENGINES
    • F01C—ROTARY-PISTON OR OSCILLATING-PISTON MACHINES OR ENGINES
    • F01C1/00—Rotary-piston machines or engines
    • F01C1/08—Rotary-piston machines or engines of intermeshing engagement type, i.e. with engagement of co- operating members similar to that of toothed gearing
    • F01C1/082—Details specially related to intermeshing engagement type machines or engines
    • F01C1/084—Toothed wheels
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2210/00—Fluid
    • F04C2210/24—Fluid mixed, e.g. two-phase fluid
    • F04C2210/242—Steam
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2230/00—Manufacture
    • F04C2230/90—Improving properties of machine parts
    • F04C2230/91—Coating
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2240/00—Components
    • F04C2240/50—Bearings
    • F04C2240/52—Bearings for assemblies with supports on both sides
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2240/00—Components
    • F04C2240/50—Bearings
    • F04C2240/54—Hydrostatic or hydrodynamic bearing assemblies specially adapted for rotary positive displacement pumps or compressors
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F05—INDEXING SCHEMES RELATING TO ENGINES OR PUMPS IN VARIOUS SUBCLASSES OF CLASSES F01-F04
    • F05C—INDEXING SCHEME RELATING TO MATERIALS, MATERIAL PROPERTIES OR MATERIAL CHARACTERISTICS FOR MACHINES, ENGINES OR PUMPS OTHER THAN NON-POSITIVE-DISPLACEMENT MACHINES OR ENGINES
    • F05C2251/00—Material properties
    • F05C2251/14—Self lubricating materials; Solid lubricants
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C17/00—Sliding-contact bearings for exclusively rotary movement
    • F16C17/02—Sliding-contact bearings for exclusively rotary movement for radial load only
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16C—SHAFTS; FLEXIBLE SHAFTS; ELEMENTS OR CRANKSHAFT MECHANISMS; ROTARY BODIES OTHER THAN GEARING ELEMENTS; BEARINGS
    • F16C17/00—Sliding-contact bearings for exclusively rotary movement
    • F16C17/12—Sliding-contact bearings for exclusively rotary movement characterised by features not related to the direction of the load
    • F16C17/14—Sliding-contact bearings for exclusively rotary movement characterised by features not related to the direction of the load specially adapted for operating in water

Definitions

  • This invention relates to the lubrication of screw machines such as screw expanders, which may be used to generate power using, for example, steam as the working fluid.
  • Positive-displacement expanders are increasingly popular for use in power generation.
  • One of the most successful positive-displacement machines is the plural-screw machine, most commonly embodied as a twin-screw machine.
  • Such machines are disclosed in UK Patent Nos. GB 1197432, GB 1503488 and GB 2092676 to Svenska Rotor Maskiner (SRM).
  • Screw machines can be used as compressors or as expanders.
  • the broadest concept of this invention is concerned with both types of screw machines, but the invention has particular benefits in the context of expanders. This specification is therefore primarily concerned with, and describes the invention in relation to, the use of screw machines as expanders. Screw machines for use as expanders will be referred to in this specification simply as screw expanders.
  • a key advantage of a screw expander over a turbine expander is the ability to handle a wet working fluid (i.e. a fluid containing both gaseous and liquid phases) with little risk of damage. This is because the fluid velocities within screw machines are approximately an order of magnitude lower than the fluid velocities typically encountered in turbine machines.
  • screw expanders can admit fluids of any composition from pure liquid to dry vapour, while maintaining thermodynamic equilibrium between the phases. Turbine expanders, in contrast, are vulnerable to blade erosion if any significant fraction of liquid phase is entrained in the working fluid.
  • Screw expanders comprise a casing having at least two intersecting bores.
  • the bores accommodate respective meshing helical lobed rotors, which contra-rotate within the fixed casing.
  • the casing encloses the rotors totally, in an extremely close fit.
  • the central longitudinal axes of the bores are coplanar in pairs and are usually parallel.
  • a male (or ‘main’) rotor and a female (or ‘gate’) rotor are mounted to the casing on bearings for rotation about their respective axes, each of which coincides with a respective one of the bore axes in the casing.
  • the rotors are normally made of metal such as mild steel but they may be made of high-speed steel. It is also possible for the rotors to be made of ceramic materials. Normally, if of metal, they are machined but alternatively they can be ground or cast.
  • the rotors each have helical lands, which mesh with helical grooves between the lands of at least one other rotor.
  • the meshing rotors effectively form one or more pairs of helical gear wheels, with their lobes acting as teeth.
  • the or each male rotor has a set of lobes corresponding to the lands and projecting outwardly from its pitch circle.
  • the or each female rotor has a set of depressions extending inwardly from its pitch circle and corresponding to the grooves of the female rotor(s).
  • the number of lands and grooves of the male rotor(s) is different to the number of lands and grooves of the female rotor(s).
  • the principle of operation of a screw expander is based on volumetric changes in three dimensions.
  • the space between any two successive lobes of each rotor and the surrounding casing forms a separate working chamber.
  • the volume of this chamber varies as rotation proceeds due to displacement of the line of contact between the two rotors.
  • the volume of the chamber is a maximum where the entire length between the lobes is unobstructed by meshing contact between the rotors.
  • the volume of the chamber is a minimum, with a value of nearly zero, where there is full meshing contact between the rotors at the end face.
  • Fluid to be expanded enters the screw expander through an opening that forms a high-pressure or inlet port, situated mainly in a front plane of the casing.
  • the fluid thus admitted fills the chambers defined between the lobes.
  • the trapped volume in each chamber increases as rotation proceeds and the contact line between the rotors recedes.
  • the filling or admission process terminates and further rotation causes the fluid to expand as it moves downstream through the screw expander.
  • a low-pressure or discharge port in the casing is exposed. That port opens further as further rotation reduces the volume of fluid trapped between the lobes and the casing. This causes the fluid to be discharged through the discharge port at approximately constant pressure. The process continues until the trapped volume is reduced to virtually zero and substantially all of the fluid trapped between the lobes has been expelled.
  • the meshing action of the lobes is essentially the same as that of helical gears.
  • the shape, of the lobes must be such that at any contact position, a sealing line is formed between the rotors and between the rotors and the casing in order to prevent internal leakage between successive chambers.
  • the chambers between the lobes should be as large as possible, in order to maximise fluid displacement per revolution.
  • the contact forces between the rotors should be low in order to minimise internal friction losses and to minimise wear.
  • the rotor profile is the most important feature in determining the flow rate and efficiency of a screw expander.
  • the earliest screw expanders used a very simple symmetric rotor profile, as shown in FIG. 1( a ).
  • the male rotor 10 comprises part-circular lobes 12 equi-angularly spaced around the pitch circle, whose centres of radius are positioned on the pitch circle 14 .
  • the profile of the female rotor 16 simply mirrors this with an equivalent set of part-circular depressions 18 .
  • Symmetric rotor profiles such as this have a very large blow-hole area, which creates significant internal leakage. This excludes symmetric rotor profiles from any applications involving a high pressure ratio or even a moderate pressure ratio.
  • SRM introduced its ‘A’ profile, shown in FIG. 1( b ) and disclosed in various forms in the aforementioned UK Patent Nos. GB 1197432, GB 1503488 and GB 2092676.
  • the ‘A’ profile greatly reduced internal leakage and thereby enabled screw compressors to attain efficiencies of the same order as reciprocating machines.
  • the Cyclon profile shown in FIG. 1( c ) reduced leakage even further but at the expense of weakening the lobes of the female rotors 16 . This risks distortion of the female rotors 16 at high pressure differences, and makes them difficult to manufacture.
  • the Hyper profile shown in FIG. 1( d ) attempted to overcome this by strengthening the female rotors 16 .
  • the ‘N’ rotor profile is characterised in that, as seen in cross section, the profiles of at least those parts of the lobes projecting outwardly of the pitch circle of the male rotor(s) and the profiles of at least the depressions extending inwardly of the pitch circle of the female rotor(s) are generated by the same rack formation.
  • the latter is curved in one direction about the axis of the male rotor(s) and in the opposite direction about the axis of the female rotor(s), the portion of the rack which generates the higher pressure flanks of the rotors being generated by rotor conjugate action between the rotors.
  • a portion of the rack preferably that portion which forms the higher pressure flanks of the rotor lobes, has the shape of a cycloid.
  • the bottoms of the grooves of the male rotor(s) lie inwardly of the pitch circle as ‘dedendum’ portions and the tips of the lands of the female rotor(s) extend outwardly of its pitch circle as ‘addendum’ portions.
  • these dedendum and addendum portions are also generated by the rack formation.
  • the pitch circles P have radii proportional to the number of lands and grooves on the respective rotors.
  • ⁇ is the rotation angle of the main rotor for which the primary and secondary arcs have a contact point. This angle meets the conjugate condition described by Sakun in Vintovie kompressori, Mashgiz Leningrad, 1960:
  • a special coordinate system of this type is a rack (rotor of infinite radius) coordinate system, indicated at R in FIG. 2( b ), which shows one unit of a rack for generating the profiles of the rotors shown in FIG. 2( a ).
  • An arc on the rack is then defined as an arbitrary function of a parameter:
  • ⁇ represents a rotation angle of the rotor where a given arc is projected, defining a contact point. This angle satisfies the condition (5) which is:
  • FIG. 2( c ) shows the relationship of the rack formation of FIG. 2( b ) to the rotors shown in FIG. 2( a ), and shows the rack and rotors generated by the rack.
  • FIG. 2( d ) shows the outlines of the rotors shown in FIG. 2( c ) superimposed on a prior art rotor pair by way of comparison.
  • rack generation offers two advantages compared with rotor coordinate systems: a) a rack profile represents the shortest contact path in comparison with other rotors, which means that points from the rack will be projected onto the rotors without any overlaps or other imperfections; b) a straight line on the rack will be projected onto the rotors as involutes.
  • the profile is usually produced by a conjugate action of both rotors, which undercuts the high pressure side of them.
  • the practice is widely used: in GB 1197432, singular points on main and gate rotors are used; in GB 2092676 and GB 2112460 circles were used; in GB 2106186 ellipses were used; and in EP 0166531 parabolae were used.
  • An appropriate undercut was not previously achievable directly from a rack. It was found that there exists only one analytical curve on a rack which can exactly replace the conjugate action of rotors.
  • This is preferably a cycloid, which is undercut as an epicycloid on the main rotor and as a hypocycloid on the gate rotor. This is in contrast to the undercut produced by singular points which produces epicycloids on both rotors. The deficiency of this is usually minimized by a considerable reduction in the outer diameter of the gate rotor within its pitch circle. This reduces the blow-hole area, but also reduces the throughput.
  • a conjugate action is a process when a point (or points on a curve) on one rotor during a rotation cuts its (or their) path(s) on another rotor.
  • An undercut occurs if there exist two or more common contact points at the same time, which produces ‘pockets’ in the profile. It usually happens if small curve portions (or a point) generate long curve portions, when considerable sliding occurs.
  • the ‘N’ rotor profile overcomes this deficiency because the high pressure part of a rack is generated by a rotor conjugate action which undercuts an appropriate curve on the rack. This rack is later used for the profiling of both the main and gate rotors by the usual rack generation procedure.
  • the coordinates of all primary arcs on the rack are summarised here relative to the rack coordinate system.
  • the lobe of this profile is divided into several arcs.
  • the rack coordinates are obtained through the procedure inverse to equations (7) to (11).
  • FIG. 2( d ) shows the profiles of main and gate rotors 3 , 4 generated by this rack procedure superimposed on the well-known profiles 5 , 6 of corresponding rotors generated in accordance with GB 2092676, in 5/7 configuration.
  • the rack-generated profiles give an increase in displacement of 2.7% while the lobes of the female rotor are thicker and thus stronger.
  • the segments AB, BC, CD, DE, EF and FG are all generated by equation (12) above.
  • the values of p and q may vary by ⁇ 10%.
  • the segments BC, DE and FG r is greater than the pitch circle radius of the main rotor, and is preferably infinite so that each such segment is a straight line.
  • the ‘N’ rotor profile described above is based on the mathematical theory of gearing.
  • the relative motion between the rotors is very nearly pure rolling: the contact band between the rotors lies very close to their pitch circles.
  • the ‘N’ rotor profile has many additional advantages over other rotor profiles, which include low torque transmission and hence small contact forces between the rotors, strong female rotors, large displacement and a short sealing line that results in low leakage. Overall its use raises the adiabatic efficiencies of screw expander machines, especially at lower tip speeds, where gains of up to 10% over other rotor profiles in current use have been recorded.
  • external timing gears may be omitted, such that synchronisation of the rotors is determined solely by their meshed relationship. This necessarily implies some transmission of synchronising torque from one rotor to the other via their meshed helical formations. In that case, the helical formations of the rotors must be lubricated to avoid hard contact between the rotors, with consequent wear and probable seizure.
  • An oil-flooded machine relies on oil entrained in the working fluid to lubricate the helical formations of the rotors and their bearings and to seal the gaps between the rotors and between the rotors and the surrounding casing. It requires an external shaft seal but no internal seals and is simple in mechanical design. Consequently, it is cheap to manufacture, compact and highly efficient.
  • an oil-free machine does not mix the working fluid with oil.
  • timing gears are provided to avoid contact between the helical formations of the rotors.
  • Each timing gear wheel turns with a respective one of the rotors and those gear wheels mesh outside the casing, where they are lubricated externally with oil.
  • oil-free refers to the interior of the casing rather than to the machine as a whole.
  • internal seals are required on each shaft between the casing and the gear wheels, as well as an external shaft seal. It follows that oil-free machines are considerably bulkier and much more expensive to manufacture than oil-flooded machines. However, the rotors of oil-free machines can rotate at higher speeds, without excessive viscous drag. Hence, the flow capacity per unit volume of oil-free machines is higher than for oil-flooded machines.
  • Both oil-flooded and oil-free machines require an external heat exchanger for the lubricating oil before it is readmitted to the machine.
  • the purpose of the heat exchanger differs between oil-flooded and oil-free machines.
  • An oil-free machine uses the heat exchanger to cool the oil.
  • an oil tank, oil filters and a circulating pump are required to return the oil to the bearings and the timing gears.
  • an oil-flooded machine requires a separator downstream of the expander to remove the oil from the discharged working fluid. The separated oil then has to be re-pressurised in a pump and the heat exchanger has to heat the oil before it is returned to the high-pressure end of the casing. This is to avoid chilling the working fluid entering the casing, which otherwise would reduce the efficiency of the expander.
  • oil-flooded expanders In the case of oil-flooded expanders, it is practically impossible to separate and remove the oil completely from the working fluid after expansion. This leads to a gradual accumulation of oil in other parts of the system, which creates operating problems. Of course, the same problem arises in oil-free expanders whose sealing arrangements are unable to keep the oil completely separate from the working fluid.
  • oil-free expanders are generally better in this respect than oil-flooded expanders; thus, applications that are particularly sensitive to oil contamination may necessitate the adoption of oil-free expanders despite their bulk, complexity and much higher cost.
  • the rotors of a screw expander are mounted to the casing on bearings for rotation about their respective axes.
  • bearings Various types may be used.
  • Lubrication is also an issue here of course.
  • Rolling-element bearings include ball bearings and roller bearings. They function by maintaining rolling contact through a set of spherical balls or cylindrical or frusto-conical rollers that separate two surfaces. If suitably arranged, rolling-element bearings may sustain both radial and axial loads. Despite the predominantly rolling motion between the rolling elements and the tracks on which they run, a boundary film of oil must be maintained between those parts to minimise wear and frictional heating.
  • WO 2006/131759 shows that it is possible to lubricate rolling-element bearings in a screw expander even where a liquid component of the working fluid contains only a low concentration of dissolved oil. If that liquid component is supplied to the bearings, the working fluid will evaporate due to frictional heating and will leave enough oil in the bearing housing to supply the boundary film needed to keep the bearings operating effectively.
  • the lubricating principles of WO 2006/131759 cannot be applied when using a screw expander for expanding fluids such as steam, or any other fluids in which oil cannot be dissolved in its liquid phase, or where the presence of even small traces of lubricating oil in the working fluid is not permissible.
  • U.S. Pat. No. 6,217,304 to Shaw discloses a screw compressor for refrigeration apparatus that, in theory, can use droplets of liquid refrigerant entrained in the gaseous phase of the refrigerant to seal, cool and lubricate the rotors. There is provision to inject droplets of liquid refrigerant into the refrigerant flow if necessary, with the penalty of complication.
  • U.S. Pat. No. 6,217,304 does not provide an enabling disclosure of how such a compressor may be run without oil in the working fluid. It merely refers to the use of a thermoplastic or other suitable composite material for the male rotor that allows a small clearance between the rotors.
  • the invention resides in a screw machine for use with a working fluid containing a liquid phase, the machine comprising two or more rotors having meshed, lubricated helical formations, wherein the rotors have an ‘N’ profile as defined herein and, in use, lubrication of the helical formations of the rotors is achieved exclusively—or at least substantially so—with the liquid phase of the working fluid.
  • the invention may also be expressed as a method of lubricating a screw machine when using a working fluid containing a liquid phase, the machine comprising two or more rotors having meshed, lubricated helical formations, the rotors having an ‘N’ profile as defined herein, wherein the method comprises lubricating the helical formations of the rotors substantially exclusively with the liquid phase of the working fluid.
  • references to ‘substantially exclusively’ in this specification are intended to reflect that minor or trace amounts of various other fluids could be entrained in the working fluid, even if not deliberately added to it, and could have some very slight lubricating effect. However, effective lubrication still depends wholly or predominantly on the presence of the liquid phase of the specified working fluid, such that its absence would render lubrication ineffective.
  • the lubricating liquid entrained in the working fluid is preferably derived from the working fluid entering the expander, substantially without prior addition of further liquid to the working fluid. This is an advantageously simple arrangement. However, prior addition of further liquid is possible if necessary and is not excluded from the invention in its broadest sense.
  • liquid phase of the working fluid as the lubricant for the helical formations obviates expensive lubrication systems for delivering oil lubricant to those formations. It also avoids oil contamination of the working fluid and the need for separating oil from the working fluid after the working fluid has passed through the machine.
  • the rotors may be manufactured of any suitable material.
  • their helical formations may be coated with a low-friction coating, for example Balinit C2 (trade mark) as offered by Oerlikon Balzers.
  • Balinit C2 is a ‘WC/C’ coating that is deposited by physical vapour deposition (PVD) and comprises a carbide phase and a carbon phase.
  • PVD physical vapour deposition
  • the use of a low-friction coating on at least the helical formations of the rotors is preferred where wear characteristics make such a coating desirable.
  • it is preferred that the helical formations of the rotors are left uncoated if possible; it is a potential benefit of the invention to allow this by virtue of the use of ‘N’-profile rotors.
  • the invention also resides in a screw machine for use with a working fluid containing a liquid phase, the machine comprising two or more rotors having meshed, lubricated helical formations coated with a low-friction coating, wherein, in use, lubrication of the helical formations of the rotors is achieved substantially exclusively with the liquid phase of the working fluid.
  • That aspect of the invention may also be expressed as a method of lubricating a screw machine when using a working fluid containing a liquid phase, the machine comprising two or more rotors having meshed, lubricated helical formations coated with a low-friction coating, wherein the method comprises lubricating the helical formations of the rotors substantially exclusively with the liquid phase of the working fluid.
  • the rotors are preferably supported by bearings that are also substantially exclusively lubricated, in use, by the liquid phase of the working fluid.
  • the use of the liquid phase to lubricate the bearings obviates expensive lubrication systems for delivering oil lubricant to those bearings. It also avoids oil contamination of the working fluid and, again, the need for separating oil from the working fluid after the working fluid has passed through the machine.
  • Bearings lubricated by the liquid phase of the working fluid are preferably hydrodynamic for simplicity, but they may be hydrostatic. Whilst the larger clearances required for hydrodynamic bearings and consequent whirling of the shaft within the bearing may make the machine slightly less efficient than an equivalent machine using rolling-element bearings, screw steam expanders can be produced using water in the liquid phase as the bearing lubricant.
  • Hydrodynamic bearings operate by maintaining a film of lubricant between the rotating or sliding parts and their static casing, so that there is no contact between them except at start-up and shutdown.
  • the basic principle of operation is that the film is not of uniform thickness. In the case of journal bearings, this happens where the centre of rotation of the shaft is displaced from the centre of radius of the surrounding casing. This creates a non-uniform film of lubricant around the shaft, which results in a huge increase of pressure in the lubricant in the region where the film is thinnest.
  • the difference in pressure around the film is sufficient to urge the shaft back into alignment with the surrounding casing, hence preventing the shaft and the casing from coming into contact.
  • the pressures created in the film of lubricant depend on the viscosity of the lubricant and the reduction of thickness achieved in the lubricant film. Normally such bearings are oil-lubricated. However, some specialist firms such as Waukesha Bearings of Wisconsin, USA have developed hydrodynamic bearings that employ liquids of very low viscosity, such as water and light hydrocarbons. The key to the success of these bearings is the ability to operate with very fine minimum film thicknesses and also the use of bearing materials that do not readily seize or abrade during start-up and shutdown.
  • Hydrostatic bearings are also possible within the broad inventive concept but they are less preferred as they need an external pump and circulating system to implement them. Specifically, such bearings keep the rotor shaft from making contact with the surrounding casing by a series of pads in the casing through which high-pressure gas or liquid is admitted. The balance of forces from the pressure between the rotor shaft and the pads prevents contact being made between them and the rotor shaft is supported by the fluid as it revolves.
  • Rolling element bearings such as those of the ball or roller type, when suitably designed, can also be used to support the rotors.
  • the working fluid is, most advantageously, water or wet steam such that the liquid phase used to lubricate the helical formations, and optionally also the bearings, is liquid water such as may be entrained in the flow of steam.
  • screw machines could also be similarly designed for any working fluid such as hydrocarbons or refrigerants, provided that the liquid phase of the same fluid is readily available for use in lubricating the helical formations of the rotor and optionally also the bearings.
  • a low-viscosity liquid such as water to lubricate the helical formations of the rotors has the further advantage that the lubricant gives rise to less viscous drag than would oil.
  • the rotors can rotate at higher speed than in an equivalent oil-flooded machine, to the benefit of flow capacity.
  • the rotors are not linked by separate timing gears.
  • the entire oil lubrication system and the additional components required either for oil-free or oil-flooded machines can be eliminated if the working fluid can also be used to lubricate the bearings.
  • the compressor was then run for 150 hours with contact between the helical formations of the rotors, lubricated only by water, injected into the working fluid instead of oil. At the end of that period, the rotors were examined and showed no signs of wear or damage, other than a light polish on the contact band.
  • Example 1 The coated rotors of Example 1 were replaced by a pair of ‘N’-profile rotors of uncoated steel and run for a further five hours. At the end of that period, the uncoated rotors were examined and they, too, showed no significant signs of wear.
  • FIGS. 1( a ) to 1 ( d ) and 2 ( a ) to 2 ( d ) of the accompanying drawings describe some prior art rotor profiles.
  • FIGS. 1( a ) to 1 ( d ) and 2 ( a ) to 2 ( d ) of the accompanying drawings describe some prior art rotor profiles.
  • a screw expander 20 comprises a fixed casing 22 containing two meshing helical lobed rotors 10 , 16 that contra-rotate within the casing about parallel axes.
  • the rotors 10 , 16 are of any suitable material such as steel, optionally coated with a low-friction coating such as the aforementioned Balinit C2. They have an ‘N’ rotor profile as disclosed by the Applicant in WO 97/43550.
  • Each rotor 10 , 16 is mounted to a respective shaft 24 , 26 that is mounted, in turn, to the casing 22 by hydrodynamic bearings 28 supporting each end of each shaft.
  • One of the shafts 26 extends out of the casing 22 to drive a generator (not shown) for producing electricity.
  • the working fluid to be expanded enters the casing 22 at high pressure through an inlet port 30 .
  • the steam flows and expands through the interior of the casing 22 , causing the rotors 10 , 16 to turn at high speed, and exits the casing 22 at lower pressure through a discharge port 32 .
  • the bearings 28 are lubricated by water derived from the wet steam feed.
  • a reservoir 34 communicates with the inlet port 30 and provides water under pressure through feed lines 36 to each of the bearings 28 .
  • a balance piston 38 at the end of each rotor shaft 24 , 26 uses the pressure of the working fluid to oppose axial loading of the shaft, hence reducing the axial loading experienced by the bearings 28 that support the shafts.
  • a pressure line 40 connects the balance pistons 38 to the inlet port 30 .
  • the working fluid for expansion may be obtained from various sources, such as steam from a geothermal source.
  • a key advantage of a screw expander over a turbine expander is the ability to handle a wet working fluid (i.e. a fluid containing both gaseous and liquid phases) with little risk of damage.
  • a screw expander is also much better than a turbine expander at handling contaminated or dirty working fluids, such as wet steam containing fine particles of sand from a geothermal source or rust from corroded pipework.
  • Another advantage is that screw expanders are potentially more cost-effective than turbines for relatively small power outputs.
  • each branch pipe typically has a control valve that throttles the steam to whatever lower pressure may be required for the process in question. It is possible to use a screw expander instead of a throttle valve to reduce steam pressure. This makes it possible to recover power from the expansion process while still supplying steam at the required lower pressure.
  • the benefits of cost, robustness, compactness, reliability, efficiency and avoidance of oil contamination allowed by the present invention are crucial to the acceptance of steam expanders for such applications, especially where there is scope to replace multiple throttle valves in an industrial installation.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Applications Or Details Of Rotary Compressors (AREA)
  • Sliding-Contact Bearings (AREA)
US13/578,348 2010-02-12 2011-02-14 Lubrication of screw machines Abandoned US20130052072A1 (en)

Applications Claiming Priority (3)

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GB1002411.5 2010-02-12
GB1002411.5A GB2477777B (en) 2010-02-12 2010-02-12 Lubrication of screw expanders
PCT/GB2011/050275 WO2011098835A2 (fr) 2010-02-12 2011-02-14 Lubrification de machines à vis

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US20130052072A1 true US20130052072A1 (en) 2013-02-28

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EP (1) EP2534339A2 (fr)
JP (1) JP5964245B2 (fr)
KR (1) KR20120140659A (fr)
CN (1) CN102834618B (fr)
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EP3232021A1 (fr) * 2016-04-14 2017-10-18 Kabushiki Kaisha Kobe Seiko Sho (Kobe Steel, Ltd.) Dispositif de récupération d'énergie thermique et son procédé de fonctionnement
WO2017189022A1 (fr) * 2016-04-29 2017-11-02 Imo Industries, Inc. Ensemble de rotor de compensation de poussée modulaire
US9932829B2 (en) 2013-05-31 2018-04-03 Kobe Steel, Ltd. Expander
CN108087037A (zh) * 2018-01-22 2018-05-29 中国石油大学(华东) 一种闭式双螺杆膨胀机发电装置
US20210260497A1 (en) * 2018-02-11 2021-08-26 John D. Walker Single-temperature-thermal-energy-storage
US20230294014A1 (en) * 2019-02-11 2023-09-21 John D. Walker Enhanced power and desalination performance in medx plant design utilizing brine-waste and single-temperature- thermal energy storage coupled to thermal vapor expander
US11898561B2 (en) 2019-05-20 2024-02-13 Carrier Corporation Direct drive refrigerant screw compressor with refrigerant lubricated rotors
US11959484B2 (en) 2019-05-20 2024-04-16 Carrier Corporation Direct drive refrigerant screw compressor with refrigerant lubricated bearings

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CN105829716B (zh) * 2013-12-18 2019-05-31 开利公司 提高压缩机轴承可靠性的方法
DE102014010149B3 (de) * 2014-07-03 2015-08-13 Knut Denecke Verfahren zum Verdichten eines Dampfes und Dampfverdichter
KR101996400B1 (ko) * 2016-06-23 2019-07-04 한국조선해양 주식회사 증발가스 재액화 시스템 및 선박
DE102017121954A1 (de) * 2017-09-21 2019-03-21 GasNet s.r.o. Schraubenexpander und Verfahren zum Erzeugen von mechanischer Energie durch Expandieren eines Arbeitsfluids
GB2578923B (en) * 2018-11-14 2021-05-26 Edwards Ltd A rotor for a twin shaft pump and a twin shaft pump
CN109356659B (zh) * 2018-12-25 2024-01-02 中国石油大学(华东) 一种双螺杆膨胀机的锥形螺杆转子
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Cited By (11)

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Publication number Priority date Publication date Assignee Title
US20140199200A1 (en) * 2013-01-14 2014-07-17 Magna Powertrain Ag & Co Kg Expander circuit
US9932829B2 (en) 2013-05-31 2018-04-03 Kobe Steel, Ltd. Expander
EP3232021A1 (fr) * 2016-04-14 2017-10-18 Kabushiki Kaisha Kobe Seiko Sho (Kobe Steel, Ltd.) Dispositif de récupération d'énergie thermique et son procédé de fonctionnement
CN107299844A (zh) * 2016-04-14 2017-10-27 株式会社神户制钢所 热能回收装置及其运转方法
WO2017189022A1 (fr) * 2016-04-29 2017-11-02 Imo Industries, Inc. Ensemble de rotor de compensation de poussée modulaire
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CN108087037A (zh) * 2018-01-22 2018-05-29 中国石油大学(华东) 一种闭式双螺杆膨胀机发电装置
US20210260497A1 (en) * 2018-02-11 2021-08-26 John D. Walker Single-temperature-thermal-energy-storage
US20230294014A1 (en) * 2019-02-11 2023-09-21 John D. Walker Enhanced power and desalination performance in medx plant design utilizing brine-waste and single-temperature- thermal energy storage coupled to thermal vapor expander
US11898561B2 (en) 2019-05-20 2024-02-13 Carrier Corporation Direct drive refrigerant screw compressor with refrigerant lubricated rotors
US11959484B2 (en) 2019-05-20 2024-04-16 Carrier Corporation Direct drive refrigerant screw compressor with refrigerant lubricated bearings

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GB2477777B (en) 2012-05-23
CN102834618B (zh) 2016-08-10
CN102834618A (zh) 2012-12-19
GB2477777A (en) 2011-08-17
WO2011098835A2 (fr) 2011-08-18
GB201002411D0 (en) 2010-03-31
JP2013519820A (ja) 2013-05-30
WO2011098835A3 (fr) 2012-09-27
KR20120140659A (ko) 2012-12-31
JP5964245B2 (ja) 2016-08-03
EP2534339A2 (fr) 2012-12-19

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