HDWRV Databook

HOWDEN COMPRESSORS

WRV DATABOOK

WRV COMPRESSOR DATABOOK Section

Page

1

INTRODUCTION TO THE WRV COMPRESSOR RANGE

1.1

Product Overview.......................................................................................................

1-1

1.2

Basic Compressor Identification ................................................................................

1-3

1.3

The WRV Compressor Range ...................................................................................

1-4

2

SPECIFICATION OF HOWDEN WRV COMPRESSOR

2.1

General ......................................................................................................................

2-1

2.2

Casings......................................................................................................................

2-2

2.3

Rotors ........................................................................................................................

2-3

2.4

Bearings.....................................................................................................................

2-4

2.5

Shaft Seals ................................................................................................................

2-5

2.6

Capacity Control ........................................................................................................ 2-13

2.7

Variable Vi Control ..................................................................................................... 2-15

2.8

Standard Materials of Construction ........................................................................... 2-20

2.9

Compressor Identification .......................................................................................... 2-21

2.10

Quality Assurance of WRV Compressors .................................................................. 2-23

2.11

API 619 Comments………………………………………………………………………… 2-24

3

TECHNICAL INFORMATION

3.1

Compressor Capacity and Design Limitations ...........................................................

3.2

Part Load Performance.............................................................................................. 3-42

3.3

Compressor Weights ................................................................................................. 3-45

3.4

Compressor Rotor Inertias and Starting Torque ........................................................ 3-46

3.5

Compressor Arrangement Drawings.......................................................................... 3-49

3.6

WRV Compressor Allowable Nozzle Loadings .......................................................... 3-50

3.7

Estimated Noise Levels ............................................................................................. 3-52

3.8

Typical P & I Diagram ................................................................................................ 3-58

February 2003

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3-1


Page

4

LUBRICATION OIL SYSTEMS

4.1

General ......................................................................................................................

4-1

4.2

Oil Filtration................................................................................................................

4-3

4.3

Oil Piping ...................................................................................................................

4-4

4.4

Recommended Lubricating Oils.................................................................................

4-5

5

SUPERFEED SYSTEMS

5.1

Principle of the Howden Superfeed System ..............................................................

5-1

5.2

Application of Superfeed to Refrigeration ..................................................................

5-2

5.3

Detail Design Notes on Superfeed Applications ........................................................

5-5

5.4

Superfeed System Options ........................................................................................

5-7

6

COOLING SYSTEMS

6.1

Cooling System Options ............................................................................................

6-1

6.2

External Oil Cooling ...................................................................................................

6-3

6.3

Cooling Using Liquid Refrigerant Injection.................................................................

6-4

7

CONTROL SYSTEM

7.1

Control Philosophy.....................................................................................................

7-1

7.2

Two Stage Compression Systems.............................................................................

7-2

7.3

Capacity Control ........................................................................................................

7-3

February 2003

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Page

8

SPECIFICATION OF ANCILLIARIES

8.1

Compressor Drive Coupling.......................................................................................

8-1

8.2

Gas Suction Strainer..................................................................................................

8-2

8.3

Compressor Suction and Discharge Non-Return (Check) Valves .............................

8-3

8.4

Discharge Pipe Connecting Compressor to Oil Tank/Separator................................

8-4

8.5

Oil Tank/Separator.....................................................................................................

8-5

8.6

Coalescing Oil Separators ......................................................................................... 8-18

8.7

Oil Heater................................................................................................................... 8-22

8.8

Pressure Relief Valve ................................................................................................ 8-23

8.9

Compressor By-Pass ................................................................................................. 8-24

8.10

Oil Pump Suction ....................................................................................................... 8-25

8.11

Oil Pump .................................................................................................................... 8-26

8.12

Oil Cooler................................................................................................................... 8-27

8.13

Oil Filter ..................................................................................................................... 8-28

8.14

Oil Manifold................................................................................................................ 8-29

8.15

Instrumentation .......................................................................................................... 8-30

8.16

Safety Trips................................................................................................................ 8-31

8.17

Instrument Piping....................................................................................................... 8-32

8.18

Driver Requirements. ................................................................................................. 8-33

9

SPECIAL INSTRUCTIONS FOR WRVT510 COMPRESSOR PACKAGING ..........

10

COMPUTER SELECTION PROGRAM…………………………………………………. 10-1

February 2003

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9-1

WRV COMPRESSOR DATABOOK

1

INTRODUCTION TO THE WRV COMPRESSOR PRODUCT LINE RANGE

1.1

Product Overview The WRV compressor concept is a design base from which world supply of oil injected twin screw compressors has developed following the introduction of the range to the oil, gas and refrigeration markets more than 40 years ago. The WRV product line and manufacturing processes have been constantly updated, refined and developed over many years. The purpose of this design Data Book is to maintain an up to date record of product developments and to provide original equipment supplier packaging guidelines An extensive research development and testing program instituted by Howden Compressors Limited, ensures that the WRV compressor offers significant advantages in the following areas: •

Scope of Applications Standard compressors and their model variations have demonstrated a proven history of successful operation in many applications. These include food freezing, cold storage, chilled water or glycol, turbine fuel gas boosting, landfill gas, natural gas wellhead, LNG terminal storage, offshore, hydrocarbon vapour recovery, cryogenic and many other critical process gas and industrial refrigeration applications.

•

Refrigerants and Gases WRV compressors are capable of operation with all CFC and HCFC refrigerants, anhydrous ammonia, HD-5 propane, helium, hydrogen, natural gas, CO2 and the majority of hydrocarbon gas mixtures.

•

High Pressure Capability WRV compressors are available in various casing materials to meet continuous operating pressures ratings up to: Casing Grey Iron: 24 bar g (348 psig) Nodular Iron: 32 bar g (464 psig) Steel : 36 bar g (522 psig) NB:

•

February 2003

Compressors can be supplied at up to 45 bar g discharge pressure dependent on pressure ratio.

American Petroleum Institute API 619 All compressors in the product line are available in cast steel casing material (WRVS) as a standard factory option to API 619 specification where appropriate.

1-1


•

Plain Shell SleeveType Journal Bearings White metalled journal bearings, coupled with ease of maintenance and good oil management offer the opportunity for almost limitless radial bearing operating life.

•

Heavy Duty ANSI Class Gas Casing Connections Both inlet and outlet gas connections are minimum ANSI Class 300 Flange type. Flanges are flat-faced standard and raised face flanges on steel body.

•

Infinitely Variable Slide Flow and Vi Modulation All compressors are equipped with a hydraulically actuated modulating slide valve (reduced full load volume ratio) resulting with optimum partial flow power consumption. Minimum gas turndown to 15% approaching 1.0 volume index (Vi) providing reduced starting torque profiles.

•

Adjustable Full Load Volume Index Adjustable Vi is in production on 255, 321 and 365 diameters. The adjustable Volume Index offers flexibility to apply the same model compressor in an expanded range of design suction pressures at optimum full load efficiency.

•

Economiser Side Port Vapour Injection (Super –Feed) The super-feed port is radial fed into the female rotor at the optimised location to accommodate both sub-cooling flash loads and additional evaporator side loads in refrigeration applications for maximum energy efficiency ratio (EER).

•

Natural Gas Engine Drive Capability WRV Compressors rotate clockwise when facing compressor drive shaft allowing for conventional direct-coupled engine drives.

•

New WRVi 365 Models Operational data for the larger displacement WRVi 365 models is now included in this Data Book.

•

Standard Factory Engineered Options A range of “multiple” drive shaft seals can be engineered for each WRV compressor to eliminate/control flammable or toxic gas emissions in sever environments. Sour gas contruction features are available per specifications and/or duty requirements. Power take off (PTO) features can be custom engineered for specific needs.

•

Proven Reliability With over 20,000 units operational worldwide, the WRV compressor is renowned for its operational reliability and field service capability.

February 2003

1-2


1.2

Basic Compressor Identification Consider two compressors designated 1) WRV204/16536 and 2) WRVi255/193. The coding is derived as follows:-

1)

W R V 204 165 36

-

Wet, i.e. oil injected Refrigeration or Natural Gas Volume, i.e. capacity control Rotor diameter measured in millimetres Rotor length to diameter ratio i.e. 1.65 Built-in (fixed) volume ratio of the compressor i.e. 3.6

2)

W R V i 255 193

-

Wet, i.e. oil injected Refrigeration or Natural Gas Volume, i.e. capacity control Variable volume ratio ie: 2.2 - 5.0 Rotor diameter measured in mm. Rotor length to diameter ratio, ie:1.93

Because of the ability to apply WRV compressors over a wide variety of applications, many variations of the compressor are produced and theses are designated by different and extra letter codes. The above covers only the basic coding, a complete list of the identification codes is given in Section 2.8.

February 2003

1-3


1.3

The WRV Compressor Electric Motor Drive Displacements General range data: Theoretical displacement Compressor WRV163

WRV 204

WRVi 255

WRV255 WRVi 321 WRV 321 WRVi 365

WRVT 510

3

L/D Ratio

m /hr

C.F.M.

50 Hz

60 Hz

50 Hz

60 Hz

1.45

549

658

323

387

1.80

682

818

401

481

1.10

814

976

479

574

1.45

1097

1316

645

774

1.65

1221

1464

718

861

1.93

1343

1612

790

948

1.10

1590

1906

935

1121

1.30

1756

2108

1033

1240

1.45

2157

2587

1269

1522

1.65

2400

2880

1412

1694

1.93

2635

3162

1550

1860

2.20

3199

3839

1882

2258

1.32

3840

4607

2259

2710

1.65

4799

5760

2823

3388

1.93

5272

6326

3101

3721

2.20

6399

7679

3764

4517

1.45

5868

7041

3453

4144

1.65

6677

8012

3930

4716

1.93

7810

9372

4597

5516

1.32

7679

9214

4517

5420

1.65

9598

11518

5646

6775

1.93

10540

12648

6200

7440

NOTES: 1.

Displacements are based on a 2 pole electric motor drive directly coupled at 3000 (50 Hz) and 3600 rpm (60 Hz) respectively with the exception of the WRVT 510. The WRVT 510 displacement is based on a 4 pole electric motor drive directly coupled at 1500 and 1800 rpm. Swept volume displacement is proportional to drive speed.

2.

Additional models are under development. Product specifications in this manual are subject to change without notice.

February 2003

1-4


2

SPECIFICATION OF HOWDEN WRV COMPRESSOR

2.1

General The Howden WRV compressor has been specifically designed for refrigeration and gas duties where lubricated compression is acceptable. With oil injection, exceptionally high pressure ratios in one stage are possible as the oil seals the clearances between the rotors and absorbs most of the heat of compression. The oil can be effectively separated so as to give virtually oil free delivery. As there is no need to separate the bearing housings from the compression chamber by oil and gas seals, the bearings are located close to the rotors. Therefore rotor deflections are kept to a minimum and high pressure differences across the compressor are possible. Only the power input shaft which is running at a comparatively low speed, has to be sealed to the atmosphere and a face type mechanical seal can be effectively employed for the majority of applications.

February 2003

2-1


2.2

Casings The compressor main casing is a double wall design with the outer wall of circular section. The inlet casing is dowelled and bolted to the main casing to ensure correct drive train alignment. The main casing incorporates the suction flange at the top and the discharge flange at the bottom and these are tapped to suit set pins. Due to its size the WRVT510 compressor has a separate rotor casing and a discharge casing which incorporates the side discharge flange. The standard production casings are manufactured from flake graphite cast iron to EN 1561-6GJL-250.S with minimum tensile stress of 275 N/mm2 (Comparable to ASTM A278 Grade 40B). The WRVT510 casings are manufactured from nodular (SG) iron to EN 1563-GJS-400-15.S (Comparable to ASTM A536 Grade 60/40/18). Cast steel casings can be supplied to EN 10213-3 Grade G20 Mn5 (Comparable to ASTM A352 LCC) as an option to –30O C (-22 O). Low temperature certification is also available as an option to -50 OC. (-58OF). Optional cast steel materials are available to down to -75 OC (-108OF). All pressure containment casings are hydro statically pressure tested to a minimum of 42 bar gage (609 psig) prior to assembly. Casings are mated vertically with grooved ORing static seals. Gaskets are not permitted on casing splits. SUCTION FLANGE

DOUBLE WALL MAIN CASING

OUTLET END COVER

INLET END COVER

February 2003

DISCHARGE FLANGE

2-2


2.3

Rotors The male and female rotors are located side by side in the compressor casing. They are accurately machined from solid steel bar or forgings. The male (drive) rotor consists of 4 lobes and the female rotor consists of 6 flutes. The shaft seal and journal bearing surface areas are treated and precision ground to give long life characteristics under adverse conditions. The finished rotors remain a solid steel component with no sleeves and no added welded components. Cast rotors are not permitted. The rotors have asymmetric profiles with specific profiles selected according to rotor diameter to give manimum energy efficiency ratio (EER) and co-efficient of performance (COP). The male and female rotors are dynamically balanced to ISO 1940 Grade 2.5. Standard rotor material is given in Section 2.7. For special applications rotors in other grades of steel can be supplied as an option.

February 2003

2-3


2.4

Bearings Radial bearings are steel backed 1 or 3 scallop white metal lined sleeve bearings. The journal bearings also act as seals as the oil is fed to them at a pressure higher than compressor discharge. The hydrodynamic radial bearings are replaceable precision bored steel-backed babbited shell type. Bearings are secured to prevent rotation on axial displacement. Each rotor also has a pair of steel cage angular contact anti-friction thrust bearings fitted face to face at the discharge end offering complete stability for forward and reverse thrust. Thrust bearings are off-loaded by a labyrinth type hydraulically fed dynamic balance piston on both male and female rotors. Further off-loading of the male rotor thrust on some models is achieved by having another balance piston incorporated in the input shaft seal arrangement. This balance piston by itself, gives sufficient off-loading for the WRV163 compressor and no other balance pistons are fitted to this compressor. The WRVT510 compressor has tilting pad type thrust bearings as standard. This type of thrust bearing is available as an option on WRV(i)255 and WRV(i)321 compressors. If the compressed gas contains H2S1, copper free white metal bearings are available as an option to comply with NACE Standard.

February 2003

2-4


BALANCE PISTONS INLET END RADIAL BEARINGS THRUST BEARINGS

OUTLET END RADIAL BEARINGS

BALANCE PISTON

2.5

Shaft Seals

2.5.1

Standard Shaft Seal Howden, in conjunction with a leading shaft seal manufacturer, has developed a balanced seal with a unique mounting and lubrication arrangement which ensures positive sealing and lubrication of the sealing seat with minimal carbon face oil seepage under all operating conditions. The life of the seal is extended considerably as a result of this design. The mechanical shaft seal cavity is flooded with oil as the buffer fluid. The mechanical shaft seal is mounted on the male drive rotor shaft at the suction end of the compressor and comprises a spring loaded carbon face rotating against a stationary cast iron seat. The seal is easily replaced in-situ with a minimum of disturbance to other components.

February 2003

2-5


LABYRINTH SEAL/ BALANCE PISTON

MECHANICAL SEAL SEAL SEAT

February 2003

2-6


2.5.1

Standard Shaft Seal (Continued) Minimum Distance between Drive Shaft and Compressor Shaft to remove Seal Housing Compressor Rotor Diameter 163 MM 204MM 255MM 321MM 365MM 510MM

MM 100 100 125 150 150 200

IN 4.0 4.0 5.0 6.0 6.0 8.0

Note: Distance between Motor Drive Shaft and Compressor Shaft may vary due to customers’ coupling hubs being used.

Typical Shaft Seal Seepage Rates (Drops per Minute) Compressor Rotor Diameter 163MM 204MM 255MM 321MM 365MM 510MM

Before 200 hr Break-in 3-10 4-12 5-15 6-18 6-18 6-18

After 200 hrs 0.5-2 0.5-3 1-4 2-5 2-5 2-5

Seal seapage rates are affected by shaft speed, coupling alignment, oil type and operating pressures. Above values are typical only.

February 2003

2-7


2.5.2

Double Shaft Seal Options WRV compressors can be fitted with “Double” shaft seals to eliminate/control gas emissions. A range of engineered options is available for the Double Shaft Seals that are fitted along with the Standard Shaft Seal. The following options of Double Shaft Seal arrangements are available for all WRV compressors: Fail Safe – Wet and Dry Safe T Seal This is the lowest category “Secondary Containment Seal” where only the function of outboard seal is to contain medium/oil on emergency shut down. Arrangement consists of Standard inner seal with T type outer seal which energises at approximately 0.5 bar G (7 PSIG) and deforms to act as a static seal. Seal is expendable and must be replaced after emergency trip.

February 2003

2-8


2.5.2

Double Shaft Seals (cont’d) Dual Seal Wet and Dry 28SC This is classified as medium category “Second Containment Seal” where the function can be either fail safe on shut down or combined fail safe/emission reduction system. Seal arrangement consists of Standard inner seal with type 28SC dry running dynamic outer seal. The outer seal must not be subjected to pressures higher than 3 bar G (43 PSIG).

February 2003

2-9


2.5.2

Double Shaft Seals (cont’d) Dual Seal Wet and Dry 28SC This is classified as medium category “Second Containment Seal” where the function can be either fail safe on shut down or combined fail safe/emission reduction system. Seal arrangement consists of Standard inner seal with type 28SC running dynamic outer seal. The outer seal must not be subjected to flare or purge pressures higher than 3 bar G (43 PSIG)

February 2003

2-10


2.5.2

Double Shaft Seals (cont’d) Back to Back Wet and Wet Cartridge Type This is the highest type of “Secondary Containment Seal” required for zero emission of operating gas and oil containment by gas. It also acts as a fail safe arrangement on shut down. Both inner and outer seals are Standard seals in a cartridge arrangement to facilitate assembly and site replacement.

*1 *2 *3

February 2003

CIRCULATING PUMP SELECTED FOR 1-2 bar (15-30 psi) OIL HEATER REQUIRED WHERE AMBIENT/INDOOR TEMPERATURE IS LOWER THAT 15C (60 F) LUBE OIL SKID REQUIRED

2-11


2.5.2

Double Shaft Seals (cont’d) Tandem Seal Wet and Wet This system is used on liquid pumps where bearing and seal operate on pumped medium. When used on a screw compressor it is classified as a medium category with both fail safe and emission reduction features. It is more expensive than the dual arrangement due to the extra lube oil skid and, as a result of extra equipment plus the direct contact outer seal, is more prone to damage. Both inner and outer seals are Standard design.

February 2003

2-12


2.6

Capacity Control Capacity control is achieved using a capacity control slide valve. This valve is in effect a movable section of the rotor bore casing which prevents compression beginning within the rotor lobes until the required flow is matched by the gas within the compression space. In this way, work is not done on all of the gas within the rotor lobes, hence power is saved when the system demand is less than the maximum capacity of the compressor. Capacity regulation is in the range of 100% to approximately 10% of full load. Operation at part load gives a related power saving. Graph on following page shows a typical power/capacity curve for varying condensing conditions for a water chilling unit. Table 3.2 in Section 3 shows the approximate absorbed power for different conditions of load when operating on Ammonia, Natural gas or CFC Refrigerants. The capacity control slide valve is operated using an integral hydraulic cylinder and double acting piston. To move the slide valve, oil from the lubrication system is fed, via solenoid valves, to one side or other of the piston which is connected to the slide valve. See Section 8 for connection details.

HYDRAULIC CYLINDER

Variable Vi Adjustment PISTON

Slide Valve

February 2003

CONNECTING ROD

2-13


TYPICAL PART LOAD PERFORMANCE - R22 IN COMPLETE PACKAGE SUCH AS WATER CHILLING UNIT REFRIGERATION CAPACITY BASED ON CHILLED WATER TEMPERATURE MAINTAINED AT 44°F (6.7°C) AT ALL PART LOAD CONDITIONS CONDENSING WATER TEMPERATURE 85°F (29.4°C) AT 100% RATING AND VARYING AS PER ARI STANDARD 550-92 FOR PART LOAD DUTY SIMILAR RESULTS ARE ACHIEVED WITH OTHER REFRIGERANTS UNDER SIMILAR OPERATING CONDITIONS

February 2003

2-14


2.7

Variable Vi Control

2.7.1

Principle of the variable volume ratio characteristic Unlike other types of positive displacement machines, screw compressors compress gas by volume reduction within the compressor casing before the gas reaches the discharge port. The male rotor rolls into the female flutes reducing the trapped volume. The Vi is the timing of when the gas is allowed to discharge. Volume reduction continues further internally with higher Vi set points at full load slide valve position. As the compressor fully unloads, the Vi reduces to theoretical 1.0 volume index. Variable Vi is effective at full load for peak efficiency with applications that require a wide range of compression ratio. In refrigeration, full load Vi adjustments are not necessary at lower condensing temperatures if the compressor is unloaded to match the design capacity. If wide ranges of suction and discharge conditions are required at full load, the adjustable Vi offers peak efficiency. For best full load efficiency, it is important that the degree of compression carried out between the rotor lobes closely matches that required by the compressor suction and discharge pressure conditions. If the full load internal pressure just prior to discharge is greater than that required by the system, then over compression of the gas in the discharge port would occur significantly reducing the efficiency of the compressor. If the full load internal pressure just prior to discharge is less than required, the gas discharges with some effect on power consumption. Consult the compressor selection program for quantative analysis. Theoretically, maximum full load efficiency is achieved when the compression ration within the compressor matches the ratio of the suction and discharge pressures. This may be expressed as:

Vi

∝

( Vo)

= Pco Pev

Where Vi = inlet volume. Vo = outlet volume. Pco = condensing pressure. Pev = evaporating pressure.

∝

= Cp/Cv is the adiabatic gas index.

In practice the efficiency of the compressor is also effected by other factors. The most efficient Vi may differ slightly from the theoretically calculated value.

February 2003

2-15


p.v. diagrams of internal compression 2.7.1

(Cont'd) Certain models in the WRV compressor range are fitted with a variable volume ratio (Vi) control. Provision of this facility allows the optimum Vi to be selected for the evaporating and condenser conditions. Adjustment can be made in the range Vi = 2.2 to Vi = 5.0. This adjustment of the Vi is advantageous in refrigeration where there is a significant change in the full load operating conditions. This may occur where night temperatures vary significantly from those during the day or changing seasons, and the additional full load compressor capacity at lower condensing is utilises. This is typical with multiple compressor installations where one compressor off loads via slide valve and the balance of compressors are maintained at 100% capacity. It is suggested that the daily mean condensing temperature be used as the basis for the setting. Re-adjustment of the Vi setting can be carried out at any time (See Section 2.6.2). The most significant power cost savings are achieved through unloading the compressor (Vi reduces at part load) when additional refrigeration capacity is not required. Full load performance may be optimised with changes in mean condensing conditions due to seasonal temperature changes and trends. This adjustment of the Vi is advantageous in natural gas well-head gas pumping when the equipment is designed over a wide range of full load suction pressures. This provides for maximum flexibility to move the equipment to other well-head locations. The control over this adjustment can be carried out by turning the adjuster to move the stop along the rotor length to obtain the optimum matched position The control over this adjustment can be carried out by turning the adjuster to move the stop along the rotor length to obtain the optimum matched position. Vi Adjustment It is recommended that the computer selection program be used to produce a table or curve showing the most advantageous Vi setting for the range of operating conditions likely to be encountered by the user. This information should be provided to the operators by the contractor taking account of the requirements for each particular application. To obtain this information, values for the operating conditions may be entered into the selection program which then calculates the best Vi setting taking account of all the variables within the compressor.

February 2003

2-16


2.7.2

Vi Slide Stop Setting The Vi slide stop actuating screw is located at the front end of the compressor and is adjusted manually to the desired Vi. NOTE:

Before making adjustment to the slide stop screw, the slide valve must be fully unloaded. From a safety aspect it is recommended that the compressor be stationary. Vi Adjusting Screw

The number of turns of the actuating screw to set the required Vi are shown on Pages 2-12, & 2-13, for the various compressor sizes. EXAMPLE 1 It is required to set a WRVi 255/165 at a Vi of 3.6 From the WRVi 255/165 Graph on Page 2-15 - the number of turns required is 8.8 STEP 1 STEP 2 STEP 3

Ensure slide valve is fully unloaded Rotate actuating screw clockwise until it locks at low Vi position (2.2 in this case) Turn actuating screw 8.8 turns anti-clockwise.

The Vi of 3.6 is now set. EXAMPLE 2 It is required to set a WRVi 321/193 at a Vi of 5.0 From the WRVi 321/193 Graph on Page 2-16 - the number of turns required is 24.5 STEP 1 STEP 2 STEP 3

Ensure slide valve is fully unloaded. Rotate actuating screw clockwise until it locks at low Vi position. Turn actuating screw 24.5 turns anti-clockwise.

The Vi of 5.0 is now set.

February 2003

2-17


2.7.3

Slide Valve Stop Settings – WRVi 255

February 2003

2-18


2.7.4

Slide Valve Stop Settings – WRVi 321

February 2003

2-19


2.8

Standard Materials of Construction COMPONENT COMPRESSOR CASINGS COVERS SLIDE VALVE SLIDE VALVE STOP HYDRAULIC CYLINDER WRVT510 CASINGS COVERS HYDRAULIC CYLINDER ROTORS (163, 204 & 255) (321 & 365) (510) HOUSINGS ETC COVER PLATES ETC

February 2003

MATERIAL

CAST IRON TO EN1561 –GJL-275.S ¡ASTM A48 GRADE – 40B SG IRON TO EN1563-GLS-400/15.S ¡ASTM A536 GRADE – 60/40/18 STEEL TO AISI GRADE C 1141 STEEL TO EN10250 – 2 GRADE C40 ¡AISI 1040 STEEL TO EN 10250 – 2 GRADE C55 ¡AISI 1055 STEEL BAR TO BS 970 080 M40 ¡AISI 1040 STEEL PLATE TO EN 10283-2 GRADE P265GH ¡AISI 1055

2-20


2.9

Compressor Indentification The basic coding used for compressor identification is given in Section 1.2. Due to its wide application, many variations of the compressor are produced and these variants are identified by different and extra letter codes. A full list of the letter codes used is given. The numeric coding in Section 1.2 does not vary. W

Stands for Wet i.e. oil injected and where used is the first letter in any oil injected compressor. The only alternative to W in an oil injected compressor is M.

M

Stands for Mirror image which is an oil injected compressor in which the direction of rotation is reversed in order to permit double ended drive. This is an alternative to W.

R

Stands for Refrigeration and is used to identify the compressors designed for handling refrigerants and gases, with fully sealed and hydraulically tested casings.

C

Stands for Conditioning and identifies a version of the standard refrigeration compressor which has no oil injection holes in the slide valve. It is used for some compressors operating on dense gases such as R12 and R22, propane etc. where the reduced oil quantity supplied to the compressor does not result in excessive discharge temperatures. The first use of this type was for air conditioning applications hence the use of this letter. The letter C is an alternative to R and the letters cannot be used together.

L

Stands for Light gases and identifies a version of the standard refrigeration compressor which has modified clearances and adjustments for compressors handling very light gases such as hydrogen and helium. The reduced clearances are necessary to give an acceptable performance. The letter L is an alternative to R and the letters cannot be used together.

V

Stands for Volume control and indicates that an integral controlling slide valve of some sort is fitted to the compressor.

B

Stands for Booster and applies to compressors which have a reduced pressure capability due to long rotor length, type of bearings used, or other limitations.

H

Stands for Higher pressure and applies to refrigeration/gas compressors which have an increased pressure capability due to modified thrust balance piston area, increased oil pressure supply, or other enhancements. The oil pressure of this standard increases from a nominal 30 psi (2 bar) to a nominal 40 psi (2.75 bar) differential pressure. This variant is suitable for discharge pressures up to 350 psig (24 bar g) on long rotor lengths and higher with shorter rotor lengths.

February 2003

2-21


2.9

Compressor Indentification (Cont'd) X

Stands for eXtra high pressure and applies to refrigeration/gas compressors which have pressure capability even higher than the "H" standard due to further internal modifications. The oil differential pressure used may be even higher than the level used with the "H" standard machine. The discharge pressure possibly varies with the suction pressure, the rotor L/D ratio and the particular gas being handled but is usually of the order of 435 psig (30 bar g) with L/D ratios of 1.65, and higher at smaller L/D's and lower at higher L/D's. SG iron casings and a special shaft seal/ seat and special rotor material to cope with higher torque values may be used. (Note: The 510 compressor has SG iron casings as standard for all variants. H and X, are alternatives and can never be used together.

T

Stands for Tilting pad thrust bearings which were introduced to enable the compressors to comply with API 619. (Note: The 510 compressor has tilting pad thrust bearings as standard). Discharge pressure limits are the same as for the WRVH compressors.

The coding letters for Gas/Refrigeration compressors described above are always used in the order given, W or M precedes R, C or L which precedes V etc. S

Stands for Steel casings. While the letter means steel casings, it does not specify what steel is used, which can vary depending on application.

N

Stands for Nodular cast iron and applies to compressors with nodular cast iron casings. This code letter is positioned as the S for Steel casings. As the WRVT 510 compressor is manufactured in this material as standard, the letter N is not used for its identification.

i

Stands for infinitely variable volume ratio that can be set between 2.2 and 5.0. This means that the volume ratio can be easily adjusted to suit compressor operating conditions.

February 2003

2-22


2.10

Quality Assurance ISO 9001 Howden Compressors Ltd are recognised as pioneers in the development of Rotary Screw Compressors and since the 1960’s have produced, at their factory in Craigton, Glasgow, a wide variety of oil free and oil injected versions of this equipment. The Company’s products are in use throughout the world, primarily in the food processing, oil, gas, petrochemical and power generation industries, and are supported by a network of sales offices and agents. The North American market is served by Howden Compressors LLC, near Philadelphia, which also offers an enhanced level of after sales product support. The Glasgow facility is the headquarters for design and production of screw compressors serving a wide variety of applications, including the compression of natural gas, hydrocarbon systems, refrigerants and chemical process gases. Site installation and commissioning is not carried out by Howden. Customer Support is available to offer advice as required. Packaging and delivery of product is provided by approved suppliers and packagers worldwide. Howden Compressors Limited enjoys the patronage of many prestigious customers and recognises that for this situation to continue, the Company must focus on responding to customer needs in developing our products and continually improving our systems. As such, Howden Compressors design and manufacturing systems and control procedures are fully accredited as complying with ISO 9001, 2000 and a series of ongoing internal and external audits are rigorously applied to ensure that these standards are maintained.

February 2003

2-23


2.11

API Comments

TECHNICAL SPECIFICATION COMMENTS AND DEVIATIONS TO API 619 THIRD EDITION JUNE 1997 HOWDEN OIL INJECTED WRV COMPRESSOR

Howden Compressors Limited comments and deviations from API 619, 3rd Edition, June 1997, in respect of oil injected type screw compressors. Major components are in materials to "British Standards" with nearest "American" equivalent specified in data sheets. The compressor generally complies with API 619 with the following deviations: 4.3.5

Standard Compressor does not fully comply as some connections on smaller compressor sizes have connections of 3/4 inch or less.

4.3.5.5

Unless stub pipes are agreed then this clause is excluded.

4.4.1

Standard Compressors comply with ISO standard for screw compressors and may not necessarily comply with Appendix - G.

4.5.1.2

Standard rotors use shaft integral with rotor body but not necessarily forgings. (Steel Bar ≤255mm diameter steel forging >255mm diameter)

4.5.1.3

We have not incorporated radial vibration probe sensing area. (Axial probe sensing area is provided).

4.7.1

A Torsional or Torsional Plus Lateral Analysis is available as an option for the complete shaft system. It is not carried out within the compressor due to the oil damping effects throughout the compressor and it is impracticable to attempt model analysis within the compressor by adding imbalance.

4.7.3.5

Non contacting radial vibration probes are not used in the oil injected, stiff shaft, slow speed screw compressor.

4.8.4.4*

When sleeve bearing type drive motors are used, coupling type must eliminate applied thrust load to the compressor.

4.10.2*

A lubrication system shall be furnished with each compressor to meet all the requirements of the compressor itself, but it cannot be used to supply external items of the equipment due to the inherent characteristics of the system necessary for this type of compressor.

4.10.3

API Standard 614 cannot be used in its entirety with this type of compressor. API 614 calls for a non-pressured lubrication tank vented to atmosphere whereas in an oil injected screw compressor of the type being considered the oil tank is maintained at compressor discharge pressure at all times and is a totally sealed system within the compressor package. Those requirements of API 614 which can be incorporated into a system, for example use of stainless steel oil piping after the oil filter, duplex components etc. are available as an option.

4.12

Nameplates to this standard available as material agreed with Purchaser. *

February 2003

Responsibility of Packager

2-24


3

TECHNICAL INFORMATION

3.1

Compressor Capacity and Design Limitations (Nominal)

Compressor Type Rotor Diameter

WRV163/145 mm

L/D Ratio

163.2

163.2

1.45

1.39

Min. Male Rotor Speed

rpm

1500

1500

Max. Male Rotor Speed

rpm

4500

4500

Min. Tip Speed

m/s

12.8

12.8

Max. Tip Speed

m/s

38.5

38.5

Theoretical Volume (Range)

cfm

161-484

156-468


m³/h

274-823

265-795

Built-in Volume Ratio Max. Pressure Difference Max. Inlet Pressure Max. Outlet Pressure Min. Inlet Temperature Max. Inlet Temperature Max. Outlet Temperature Hydraulic Test Pressure Max. Power (at max. speed)

Vi

2.1

2.6

3.6

5.0

5.8

psi

188

203

236

260

270

bar

13

14

16.3

18

18.7

psig

87

75

50

30

25

barg

6

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

470

kW

350

Max. Allowable Torque

lb.ft

550

(at Male Rotor)

Nm

746

Direction of rotation of male rotor looking from drive unit

February 2003

CLOCKWISE

3-1


3.1

Compressor Capacity and Design Limitations (Nominal) (Cont'd)


WRVH163/145 mm

L/D Ratio

163.2

163.2

1.45

1.39


rpm

1500

1500


rpm

4500

4500

Min. Tip Speed

m/s

12.8

12.8

Max. Tip Speed

m/s

38.5

38.5


cfm

161-484

156-468

274-823

265-795

Theoretical Volume (Range) Built-in Volume Ratio Max. Pressure Difference Max. Inlet Pressure Max. Outlet Pressure Min. Inlet Temperature

m³/h Vi

2.1

2.6

3.6

5.0

5.8

psi

230

260

309

340

350

bar

16

18

21.3

23.5

24.2

psig

100

80

50

30

25

barg

7

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

470

kW

350


lb.ft

550

(at Male Rotor)

Nm

746

Max. Inlet Temperature Max. Outlet Temperature Hydraulic Test Pressure Max. Power (at max. speed)


February 2003

CLOCKWISE

3-2


3.1

Compressor Capacity and Design Limitations (Nominal) (Cont'd) Compressor Type Rotor Diameter

WRV163/180 mm

L/D Ratio

163.2

163.2

1.8

1.74


rpm

1500

1500


rpm

4500

4500

Min. Tip Speed

m/s

12.8

12.8

Max. Tip Speed

m/s

38.5

38.5


cfm

200-601

195-585

341-1022

331-993


m³/h Vi

2.1

2.6

3.6

5.0

5.8

psi

188

203

236

260

270

bar

13

14

16.3

18

18.7

psig

87

75

50

30

25

barg

6

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

470

kW

350


lb.ft

550

(at Male Rotor)

Nm

746



February 2003

CLOCKWISE

3-3


3.1


WRVH163/180 mm

L/D Ratio

163.2

163.2

1.8

1.74


rpm

1500

1500


rpm

4500

4500

Min. Tip Speed

m/s

12.8

12.8

Max. Tip Speed

m/s

38.5

38.5


cfm

200-601

195-585

341-1022

331-993


m³/h Vi

2.1

2.6

3.6

5.0

5.8

psi

230

260

309

340

350

bar

16

18

21.3

23.5

24.2

psig

100

80

50

30

25

barg

7

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

470

kW

350


lb.ft

550

(at Male Rotor)

Nm

746



February 2003

CLOCKWISE

3-4


3.1


WRV204/110 mm

L/D Ratio

204

204

1.1

1.04


rpm

1500

1500


rpm

4500

4500

Min. Tip Speed

m/s

16

16

Max. Tip Speed

m/s

48.1

48.1


cfm

239-718

229-686

406-1219

389-1166

Theoretical Volume (Range) Built-in Volume Ratio Max. Pressure Difference Max. Inlet Pressure Max. Outlet Pressure Min. Inlet Temperature Max. Inlet Temperature Max. Outlet Temperature Hydraulic Test Pressure Max. Power (at max. speed)

m³/h Vi

2.1

2.6

3.6

5.0

5.8

psi

188

203

236

260

270

bar

13

14

16.3

18

18.7

psig

87

75

50

30

25

barg

6

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1028

kW

766


lb.ft

1200

(at Male Rotor)

Nm

1627


February 2003

CLOCKWISE

3-5


3.1


WRVH204/110 mm

L/D Ratio

204

204

1.1

1.04


rpm

1500

1500


rpm

4500

4500

Min. Tip Speed

m/s

16

16

Max. Tip Speed

m/s

48.1

48.1


cfm

239-718

229-686

406-1219

389-1166


m³/h Vi

2.1

2.6

3.6

5.0

5.8

psi

230

260

309

340

350

bar

16

18

21.3

23.5

24.2

psig

100

80

50

30

25

barg

7

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1028

kW

766


lb.ft

1200

(at Male Rotor)

Nm

1627


February 2003

CLOCKWISE

3-6


3.1


WRV204/145 mm

L/D Ratio

204

204

1.45

1.39


rpm

1500

1500


rpm

4500

4500

Min. Tip Speed

m/s

16

16

Max. Tip Speed

m/s

48.1

48.1


cfm

322-967

310-929

548-1643

526-1578


m³/h Vi

2.1

2.6

3.6

5.0

5.8

psi

188

203

236

260

270

bar

13

14

16.3

18

18.7

psig

87

75

50

30

25

barg

6

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1028

kW

766


lb.ft

1200

(at Male Rotor)

Nm

1627


February 2003

CLOCKWISE

3-7


3.1


WRVH204/145 mm

L/D Ratio

204

204

1.45

1.39


rpm

1500

1500


rpm

4500

4500

Min. Tip Speed

m/s

16

16

Max. Tip Speed

m/s

48.1

48.1


cfm

322-967

310-929

548-1643

526-1578


m³/h Vi

2.1

2.6

3.6

5.0

5.8

psi

230

260

309

340

350

bar

16

18

21.3

23.5

24.2

psig

100

80

50

30

25

barg

7

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1028

kW

766


lb.ft

1200

(at Male Rotor)

Nm

1627


February 2003

CLOCKWISE

3-8


3.1


WRV204/165 mm

L/D Ratio

204

204

1.65

1.59


rpm

1500

1500


rpm

4500

4500

Min. Tip Speed

m/s

16

16

Max. Tip Speed

m/s

48.1

48.1


cfm

359-1076

348-1045

610-1829

592-1775


m³/h Vi

2.1

2.6

3.6

5.0

5.8

psi

188

203

236

260

270

bar

13

14

16.3

18

18.7

psig

87

75

50

30

25

barg

6

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1028

kW

766


lb.ft

1200

(at Male Rotor)

Nm

1627


February 2003

CLOCKWISE

3-9


3.1


WRVH204/165 mm

L/D Ratio

204

204

1.65

1.59


rpm

1500

1500


rpm

4500

4500

Min. Tip Speed

m/s

16

16

Max. Tip Speed

m/s

48.1

48.1


cfm

359-1076

348-1045

610-1829

592-1775


m³/h Vi

2.1

2.6

3.6

5.0

5.8

psi

230

260

309

340

350

bar

16

18

21.3

23.5

24.2

psig

100

80

50

30

25

barg

7

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1028

kW

466


lb.ft

1200

(at Male Rotor)

Nm

1627


February 2003

CLOCKWISE

3-10


3.1


WRV204/193 mm

L/D Ratio

204

204

1.93

1.87


rpm

1500

1500


rpm

4500

4500

Min. Tip Speed

m/s

16

16

Max. Tip Speed

m/s

48.1

48.1


cfm

395-1185

388-1165

671-2013

660-1979


m³/h Vi

2.1

2.6

3.6

5.0

5.8

psi

188

203

236

260

270

bar

13

14

16.3

18

18.7

psig

87

75

50

30

25

barg

6

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1028

kW

766


lb.ft

1200

(at Male Rotor)

Nm

1627


February 2003

CLOCKWISE

3-11


3.1


WRVH204/193 mm

L/D Ratio

204

204

1.93

1.87


rpm

1500

1500


rpm

4500

4500

Min. Tip Speed

m/s

16

16

Max. Tip Speed

m/s

48.1

48.1


cfm

395-1185

388-1165


m³/h

671-2013

660-1979


Vi

2.1

2.6

3.6

5.0

5.8

psi

230

260

309

340

350

bar

16

18

21.3

23.5

24.2

psig

100

80

50

30

25

barg

7

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1028

kW

766


lb.ft

1200

(at Male Rotor)

Nm

1627


February 2003

CLOCKWISE

3-12


3.1


WRVi255/110 mm

L/D Ratio

255

255

1.1

1.04


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

467-1121

447-1072

794-1905

759-1821

Theoretical Volume (Range) Volume Ratio Max. Pressure Difference Max. Inlet Pressure Max. Outlet Pressure Min. Inlet Temperature Max. Inlet Temperature Max. Outlet Temperature Hydraulic Test Pressure Max. Power (at max. speed)

m³/h Vi

2.2

2.6

3.6

5.0

5.8 Fixed

psi

192

203

236

260

270

bar

13.2

14

16.3

18

18.7

psig

84

75

50

30

25

barg

5.8

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


CLOCKWISE

Vi is variable from 2.2 to 5.0. Intermediate points are included to display pressure limits. 5.8 Vi is non-variable.

February 2003

3-13


3.1


WRViH255/110 mm

L/D Ratio

255

255

1.1

1.04


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

467-1121

447-1072

794-1905

759-1821


m³/h Vi

2.2

2.6

3.6

5.0

5.8

psi

236

260

309

340

350

bar

16.3

18

21.3

23.5

24.2

psig

96

80

50

35

25

barg

6.6

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


CLOCKWISE


February 2003

3-14


3.1


WRVi255/130 mm

L/D Ratio

255

255

1.30

1.24


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

517-1240

504-1211

878-2106

857-2057


m³/h Vi

2.2

2.6

3.6

5.0

5.8

psi

192

203

236

260

270

bar

13.2

14

16.3

18

18.7

psig

84

75

50

30

25

barg

5.8

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


CLOCKWISE

Vi is variable from 2.2 to 5.0. Intermediate points are included to display pressure limits. 5.8Vi is non-variable.

February 2003

3-15


3.1



WRViH255/130 mm

L/D Ratio

255

255

1.30

1.24


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

517-1240

504-1211

878-2106

857-2057


m³/h Vi

2.2

2.6

3.6

5.0

5.8

psi

236

260

309

340

350

bar

16.3

18

21.3

23.5

24.2

psig

96

80

50

35

25

barg

6.6

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


CLOCKWISE


February 2003

3-16


3.1



WRVi255/145 mm

L/D Ratio

255

255

1.45

1.39


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

634-1522

609-1462


m³/h

1078-2586

1035-2484

Volume Ratio Max. Pressure Difference Max. Inlet Pressure Max. Outlet Pressure Min. Inlet Temperature Max. Inlet Temperature Max. Outlet Temperature Hydraulic Test Pressure Max. Power (at max. speed)

Vi

2.2

2.6

3.6

5.0

5.8

psi

192

203

236

260

270

bar

13.2

14

16.3

18

18.7

psig

84

75

50

30

25

barg

5.8

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


CLOCKWISE

Vi isvariable from 2.2 to 5.0. Intermediate points are included to display pressure limits. 5.8 Vi is non-variable.

February 2003

3-17


3.1


WRViH255/145 mm

L/D Ratio

255

255

1.45

1.39


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

634-1522

609-1462


m³/h

1078-2586

1035-2484


Vi

2.2

2.6

3.6

5.0

5.8

psi

236

260

309

340

350

bar

16.3

18

21.3

23.5

24.2

psig

96

80

50

30

25

barg

6.6

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


CLOCKWISE


February 2003

3-18


3.1


WRVi255/165 mm

L/D Ratio

255

255

1.65

1.59


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

706-1694

685-1644


m³/h

1199-2878

1164-2793


Vi

2.2

2.6

3.6

5.0

5.8

psi

192

203

236

260

270

bar

13.2

14

16.3

18

18.7

psig

84

75

50

30

25

barg

5.8

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


CLOCKWISE


February 2003

3-19


3.1


WRViH255/165 mm

L/D Ratio

255

255

1.65

1.59


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

706-1694

685-1644


m³/h

1199-2878

1164-2793


Vi

2.2

2.6

3.6

5.0

5.8

psi

236

260

309

340

350

bar

16.3

18

21.3

23.5

24.2

psig

96

80

50

30

25

barg

6.6

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


CLOCKWISE


February 2003

3-20


3.1


WRVi255/193 mm

L/D Ratio

255

255

1.93

1.87


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

775-1860

764-1833


m³/h

1317-3160

1298-3114


Vi

2.2

2.6

3.6

5.0

5.8

psi

192

203

236

260

270

bar

13.2

14

16.3

18

18.7

psig

84

75

50

30

25

barg

5.8

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


CLOCKWISE


February 2003

3-21


3.1


WRViH255/193 mm

L/D Ratio

255

255

1.93

1.87


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

775-1860

764-1833


m³/h

1317-3160

1298-3114


Vi

2.2

2.6

3.6

5.0

5.8

psi

236

260

309

340

350

bar

16.3

18

21.3

23.5

24.2

psig

96

80

50

30

25

barg

6.6

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


CLOCKWISE


February 2003

3-22


3.1


WRV255/220 mm

L/D Ratio

255

255

2.20

2.14


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

941-2258

922-2212


m³/h

1599-3837

1566-3759


Vi

2.1

2.6

3.6

5.0

5.8

psi

140

157

181

200

210

bar

9.7

10.8

12.5

14

14.5

psig

65

55

40

30

25

barg

4.5

3.8

2.8

2.0

1.7

psig

200

barg

13.8

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


CLOCKWISE

Built-in volume ratio is fixed at full load for 2.20 L/D

February 2003

3-23


3.1


WRVB255/220 mm

L/D Ratio

255

255

2.20

2.14


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

19.7

19.7

Max. Tip Speed

m/s

47.1

47.1


cfm

941-2258

922-2212


m³/h

1599-3837

1566-3759

Built-in Volume Ratio (Fixed)

Vi

2.1

2.6

3.6

5.0

5.8

psi

68

83

105

123

132

bar

4.7

5.7

7.2

8.5

9.1

psig

65

50

40

30

25

barg

4

3.5

2.8

2.0

1.7

Max. Pressure Difference Max. Inlet Pressure Max. Outlet Pressure Min. Inlet Temperature Max. Inlet Temperature Max. Outlet Temperature Hydraulic Test Pressure Max. Power (at max. speed)

psig

120

barg

8.3

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

1542

kW

1150


lb.ft

2250

(at Male Rotor)

Nm

3050


February 2003

CLOCKWISE

3-24


3.1


WRVi321/132 mm

L/D Ratio

321

321

1.32

1.26

1500

1500


rpm


rpm

3600

3600

Min. Tip Speed

m/s

25.21

25.21

Max. Tip Speed

m/s

60.5

60.5


cfm

1129-2710

1086-2606


m³/h

1919-4605

1845-4428


Vi

2.2

2.6

3.6

5.0

5.8

psi

192

203

236

260

270

bar

13.2

14

16.3

18

18.7

psig

84

75

50

30

25

barg

5.8

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

2741

kW

2044


lb.ft

4000

(at Male Rotor)

Nm

5424


CLOCKWISE

Vi is variable from 2.2 to 5.0. Intermediate points are included to display presssure limits. 5.8 Vi is non-variable.

February 2003

3-25


3.1


WRViH321/132 mm

L/D Ratio

321.2

321.2

1.32

1.26

1500

1500


rpm


rpm

3600

3600

Min. Tip Speed

m/s

25.21

25.21

Max. Tip Speed

m/s

60.5

60.5


cfm

1129-2710

1086-2606


m³/h

1919-4605

1845-4428


Vi

2.2

2.6

3.6

5.0

5.8

psi

236

260

309

340

350

bar

16.3

18

21.3

23.5

24.2

psig

96

80

50

30

25

barg

6.6

5.5

3.5

2

1.7

psig

350

barg

24.6

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

2741

kW

2044


lb.ft

4000

(at Male Rotor)

Nm

5424


CLOCKWISE


February 2003

3-26


3.1


WRVi321/165 mm

L/D Ratio

321.2

321.2

1.65

1.59

1500

1500


rpm


rpm

3600

3600

Min. Tip Speed

m/s

25.21

25.21

Max. Tip Speed

m/s

60.5

60.5


cfm

1412-3388

1369-3285


m³/h

2398-5756

2326-5581


Vi

2.2

2.6

3.6

5.0

5.8

psi

192

203

236

260

270

bar

13.2

14

16.3

18

18.7

psig

84

75

50

30

25

barg

5.8

5

3.5

2

1.7

psig

261

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

2741

kW

2044


lb.ft

4000

(at Male Rotor)

Nm

5424


CLOCKWISE


February 2003

3-27


3.1


WRViH321/165 mm

L/D Ratio

321.2

321.2

1.65

1.59

1500

1500


rpm


rpm

3600

3600

Min. Tip Speed

m/s

25.21

25.21

Max. Tip Speed

m/s

60.5

60.5


cfm

1412-3388

1369-3285


m³/h

2398-5756

2326-5581


Vi

2.2

2.6

3.6

5.0

5.8

psi

236

260

309

340

350

bar

16.3

18

21.3

23.5

24.2

psig

96

80

50

30

25

barg

6.6

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

2741

kW

2044


lb.ft

4000

(at Male Rotor)

Nm

5424


CLOCKWISE

Vi is variable from 2.2 to 5.0. Intermediate points are included to displaypressure limits. 5.8 Vi is non-variable.

February 2003

3-28


3.1


WRVi321/193 mm

L/D Ratio

321.2

321.2

1.93

1.87

1500

1500


rpm


rpm

3600

3600

Min. Tip Speed

m/s

25.21

25.21

Max. Tip Speed

m/s

60.5

60.5


cfm

1550-3721

1525-3661


m³/h

2634-6321

2592-6220


Vi

2.2

2.6

3.6

5.0

5.8

psi

192

203

236

260

270

bar

13.2

14

16.3

18

18.7

psig

84

75

50

30

25

barg

5.8

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

2741

kW

2044


lb.ft

4000

(at Male Rotor)

Nm

5424


CLOCKWISE


February 2003

3-29


3.1


WRViH321/193 mm

L/D Ratio

321.2

321.2

1.93

1.87

1500

1500


rpm


rpm

3600

3600

Min. Tip Speed

m/s

25.21

25.21

Max. Tip Speed

m/s

60.5

60.5


cfm

1550-3721

1525-3661


m³/h

2634-6321

2592-6220


Vi

2.2

2.6

3.6

5.0

5.8

psi

236

260

309

340

350

bar

16.3

18

21.3

23.5

24.2

psig

96

80

50

30

25

barg

6.6

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

2741

kW

2044


lb.ft

4000

(at Male Rotor)

Nm

5424


CLOCKWISE


February 2003

3-30


3.1


WRV321/220 mm

L/D Ratio

321.2

321.2

2.20

2.14

1500

1500


rpm


rpm

3600

3600

Min. Tip Speed

m/s

25.21

25.21

Max. Tip Speed

m/s

60.5

60.5


cfm

1882-4517

1841-4418


m³/h

3198-7675

3128-7507

Built-in Volume Ratio (Fixed) Max. Pressure Difference Max. Inlet Pressure Max. Outlet Pressure Min. Inlet Temperature Max. Inlet Temperature Max. Outlet Temperature Hydraulic Test Pressure Max. Power (at max. speed)

Vi

2.1

2.6

3.6

5.0

5.8

psi

140

157

181

200

210

bar

9.7

10.8

12.5

14

14.5

psig

60

50

40

30

25

barg

4

3.5

2.8

2.0

1.7

psig

200

barg

13.8

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

2741

kW

2044


lb.ft

4000

(at Male Rotor)

Nm

5424


February 2003

CLOCKWISE

3-31


3.1


WRVB321/220 mm

L/D Ratio

321.2

321.2

2.20

2.14

1500

1500


rpm


rpm

3600

3600

Min. Tip Speed

m/s

25.21

25.21

Max. Tip Speed

m/s

60.5

60.5


cfm

1882-4517

1841-4418


m³/h

3198-7675

3128-7507


Vi

2.1

2.6

3.6

5.0

5.8

psi

68

83

105

123

132

bar

4.7

5.7

7.2

8.5

9.1

psig

60

50

40

30

25

barg

4

3.5

2.8

2.0

1.7

psig

120

barg

8.3

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

2741

kW

2044


lb.ft

4000

(at Male Rotor)

Nm

5424


February 2003

CLOCKWISE

3-32


3.1


WRVi365/145 mm

L/D Ratio

365

365

1.45

1.45


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

28.6

28.6

Max. Tip Speed

m/s

68.8

68.8


cfm

1727-4144

1727-4144


m³/h

2934-7041

2934-7041


Vi

2.2

2.6

3.6

5.0

5.8

psi

188

203

236

260

270

bar

13

14

16.3

18

18.7

psig

87

75

50

30

25

barg

6

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

5814

kW

4335


lb.ft

8482

(at Male Rotor)

Nm

11500


CLOCKWISE


February 2003

3-33


3.1


WRViH365/145 mm

L/D Ratio

365

365

1.45

1.45


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

28.6

28.6

Max. Tip Speed

m/s

68.8

68.8


cfm

1727-4144

1727-4144


m³/h

2934-7041

2934-7041


Vi

2.2

2.6

3.6

5.0

5.8

psi

230

260

309

340

350

bar

16

18

21.3

23.5

24.2

psig

100

80

50

30

25

barg

7

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

5814

kW

4335


lb.ft

8482

(at Male Rotor)

Nm

11500


CLOCKWISE


February 2003

3-34


3.1


WRVi365/165 mm

L/D Ratio

365

365

1.65

1.65


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

28.6

28.6

Max. Tip Speed

m/s

68.8

68.8


cfm

1991-4716

1991-4716


m³/h

3383-8012

3383-8012


Vi

2.2

2.6

3.6

5.0

5.8

psi

188

203

236

260

270

bar

13

14

16.3

18

18.7

psig

87

75

50

30

25

barg

6

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

5814

kW

4335


lb.ft

8482

(at Male Rotor)

Nm

11500


CLOCKWISE


February 2003

3-35


3.1


WRViH365/165 mm

L/D Ratio

365

365

1.65

1.65


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

28.6

28.6

Max. Tip Speed

m/s

68.8

68.8


cfm

1991-4716

1991-4716


m³/h

3353-8012

3383-8012


Vi

2.2

2.6

3.6

5.0

5.8

psi

230

260

309

340

350

bar

16

18

21.3

23.5

24.2

psig

100

80

50

30

25

barg

7

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

5814

kW

4335


lb.ft

8482

(at Male Rotor)

Nm

11500


CLOCKWISE

Vi is variable from 2.2 to 5.0. Intermediate points are included to display pressure limits.

5.8 Vi is non-variable.

February 2003

3-36


3.1


WRVi365/193 mm

L/D Ratio

365

365

1.93

1.93


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

28.6

28.6

Max. Tip Speed

m/s

68.8

68.8


cfm

2300-5516

2300-5516


m³/h

3905-9372

3905-9372


Vi

2.2

2.6

3.6

5.0

5.8

psi

188

203

236

260

270

bar

13

14

16.3

18

18.7

psig

87

75

50

30

25

barg

6

5

3.5

2

1.7

psig

260

barg

18

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

5814

kW

4335


lb.ft

8482

(at Male Rotor)

Nm

11500


CLOCKWISE


February 2003

3-37


3.1


WRViH365/193 mm

L/D Ratio

365

365

1.93

1.93


rpm

1500

1500


rpm

3600

3600

Min. Tip Speed

m/s

28.6

28.6

Max. Tip Speed

m/s

68.8

68.8


cfm

2300-5516

2300-5516


m³/h

3905-9372

3905-9372


Vi

2.2

2.6

3.6

5.0

5.8

psi

230

260

309

340

350

bar

16

18

21.3

23.5

24.2

psig

100

80

50

30

25

barg

7

5.5

3.5

2

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

5814

kW

4335


lb.ft

8482

(at Male Rotor)

Nm

11500


CLOCKWISE


February 2003

3-38


3.1


WRVT510/132 mm

L/D Ratio

510

510

1.32

-


rpm

700

700


rpm

2000

2000

Min. Tip Speed

m/s

26.7

26.7

Max. Tip Speed

m/s

53.4

53.4


cfm

2108-6022


m³/h

3581-10231

To Contract


Vi

2.1

2.6

3.6

5.0

5.8

psi

230

264

305

336

348

bar

16

18.2

21

23.2

24.0

psig

87

75

50

31

25

barg

6

5

3.4

2.1

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

6700

kW

5000


lb.ft

18430

(at Male Rotor)

Nm

25000


February 2003

CLOCKWISE

3-39


3.1


WRVT510/165 mm

L/D Ratio

510

510

1.65

-


rpm

700

700


rpm

2000

2000

Min. Tip Speed

m/s

26.7

26.7

Max. Tip Speed

m/s

53.4

53.4


cfm

2635-7528


m³/h

4479-12789

To Contract


Vi

2.1

2.6

3.6

5.0

5.8

psi

230

264

305

336

348

bar

16

18.2

21

23.2

24.0

psig

87

75

50

31

25

barg

6

5

3.4

2.1

1.7

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

6700

kW

5000


lb.ft

18430

(at Male Rotor)

Nm

25000


February 2003

CLOCKWISE

3-40


3.1


WRVT510/193 mm

L/D Ratio

510

510

1.93

-


rpm

700

700


rpm

2000

2000

Min. Tip Speed

m/s

26.7

26.7

Max. Tip Speed

m/s

53.4

53.4


cfm

2893-8266


m³/h

4916-14044

To Contract

Built-in Volume Ratio (Fixed)

Vi

2.1

2.6

3.6

5.0

5.8

psi

230

264

305

336

348

bar

16

18.2

21

23.2

24.0

psig

87

75

50

31

25

barg

6

5

3.4

2.1

1.7

Max. Pressure Difference Max. Inlet Pressure Max. Outlet Pressure Min. Inlet Temperature Max. Inlet Temperature Max. Outlet Temperature Hydraulic Test Pressure Max. Power (at max. speed)

psig

350

barg

24.1

°F

-76

°C

-60

°F

122

°C

50

°F

212

°C

100

psig

600

barg

42

hp

6700

kW

5000


lb.ft

18430

(at Male Rotor)

Nm

25000


February 2003

CLOCKWISE

3-41


3.2

Part Load Performance % Full Load Absorbed Power %

AMMONIA

R22

Capacity

V.R.

V.R.

2.6

3.6

5.0

2.6

3.6

5.0

100

100

100

100

100

100

100

90

93

94

95

93

94

95

80

84

86

89

85

87

89

70

75

78

82

77

80

83

60

66

71

75

70

73

78

50

59

64

69

63

68

72

40

52

59

64

56

62

68

30

46

54

60

50

58

64

20

40

51

57

45

55

61

10

36

48

55

40

53

59

NOTE 1:

V.R. = Volume Ratio.

NOTE 2:

The above applies at constant condensing temperature. A considerable reduction in part load power occurs if the condensing temperature reduces under part load.

NOTE 3:

The above percentages will increase slightly at higher Pressure Ratios (PR) than those assumed above, i.e. for VR = 2.6: PR = 3, for VR = 3.6: PR = 5, for VR = 5.0: PR = 7.

February 2003

3-42


TYPICAL PART LOAD PERFORMANCE - AMMONIA ALLOWING FOR VARYING CONDENSING TEMPERATURE REFRIGERATION CAPACITY BASED ON 10°F (5.6°C) SUPERHEAT AT COMPRESSOR SUCTION AND CONDENSED LIQUID

NO

SUBCOOLING

OF

EVAPORATION TEMPERATURE: -30°F BUILT-IN VOLUME RATIO: Vi = 2.63 NOTE:

February 2003

CONDENSING TEMPERATURE VARYING FROM 95°F TO 75°F BETWEEN 100% AND 10% CAPACITY

3-43


TYPICAL PART LOAD PERFORMANCE - R22 ALLOWING FOR VARYING CONDENSING TEMPERATURE REFRIGERATION CAPACITY BASED ON 10°F (5.6°C) SUPERHEAT AT COMPRESSOR SUCTION AND CONDENSED LIQUID

NO

EVAPORATION TEMPERATURE: -30°F BUILT-IN VOLUME RATIO: Vi = 2.63 NOTE:

February 2003

CONDENSING TEMPERATURE VARYING FROM 95°F TO 75°F BETWEEN 100% AND 10% CAPACITY

3-44

SUBCOOLING

OF


3.3

Compressor Weights Compressor

WRV163

WRV204

WRVi255

WRV255

WRVi321

WRV321

WRVi365

WRVT510

February 2003

L/D

Total Compressor Weight

Ratio

kg

lb

1.45

470

1035

1.80

485

1070

1.10

760

1675

1.45

840

1850

1.65

855

1885

1.93

900

1985

1.10

1150

2535

1.30

1205

2655

1.45

1270

2800

1.65

1340

2955

1.93

1405

3095

2.20

1585

3495

1.32

2800

6170

1.65

3020

6655

1.93

3105

6845

2.20

3610

7955

1.45

4940

10890

1.65

5150

11355

1.93

5430

11970

1.32

10800

23810

1.65

11500

25353

1.93

11800

26014

3-45


3.4

Compressor Rotor Inertias and Starting Torque

Compressor

WRV163

WRV204

WRVi255

WRV255

WRVi321

WRV321

WRVi365

WRVT510

NOTE:

February 2003

L/D Ratio

Total Compressor moment of Inertia (Wr2)

Total Compressor Breakaway Torque

kg.m2

lb.ft2

kg.m

lb.ft

1.45

0.096

2.3

4.2

30

1.80

0.1168

2.8

4.9

35.5

1.10

0.1865

4.4

4.3

31

1.45

0.2929

6.9

5.6

40.5

1.65

0.3292

7.8

5.7

41

1.93

0.3800

9.0

6.5

47

1.10

0.7028

16.6

5.4

39

1.30

0.8135

19.2

6

43.5

1.45

0.8943

21.2

6.5

47

1.65

1.0047

23.8

7.2

52

1.93

1.1595

27.4

8.1

58.5

2.20

1.3086

31.0

9

65

1.32

3.0951

73.2

7.7

55.5

1.65

3.1610

74.8

9

65

1.93

3.6499

86.3

10.2

74

2.20

4.1215

97.5

11.3

82

1.45

5.2597

124.8

TBA

TBA

1.65

5.9256

140.6

TBA

TBA

1.93

6.8579

162.7

TBA

TBA

1.32

23.163

548.0

51.2

370

1.65

31.765

751.5

59.4

430

1.93

36.216

856.7

67.8

490

Inertia values are referred to male rotor input shaft.

3-46

ABSORBED TORQUE/SPEED CURVES

February 2003

VALUES SHOWN REFER TO THE MALE ROTOR INPUT SHAFT. TORQUE SHOWN IS THE ABSORBED TORQUE. AVAILABLE TORQUE MUST BE GREATER TO START COMPRESSOR.

ELECTRIC MOTOR DRIVE

WHERE A DRIVE UNIT WITH ACCELERATION MUCH LOWER THAN AN ELECTRIC MOTOR IS TO BE USED TORQUE/SPEED CURVE No. RDB 3.4.03 SHOULD BE USED.

NOTE:

VALVE AT MINIMUM CAPACITY POSITION APPLIES TO ALL VOLUME RATIOS. AVAILABLE POWER SHOULD BE 10% GREATER THAN COMPRESSOR FULL LOAD ABSORBED POWER

TYPICAL CURVES


3-47

ABSORBED TORQUE/SPEED CURVES

NOTE:

SLOW ACCELERATION DRIVE UNIT

VALUES SHOWN REFER TO THE MALE ROTOR INPUT SHAFT. TORQUE SHOWN IS THE ABSORBED TORQUE. AVAILABLE TORQUE MUST BE GREATER TO START COMPRESSOR.

VALVE AT MINIMUM CAPACITY POSITION APPLIES TO ALL VOLUME RATIOS. AVAILABLE POWER SHOULD BE 10% GREATER THAN COMPRESSOR FULL LOAD ABSORBED POWER

TYPICAL CURVES


February 2003

3-48


3.5

Summary of Latest Drawing Issue Compressor

Mk No.

External Arrgt Drg

Horizontal Sect Drg

Vertical Sect Drg

WRV (H) 163 145 & 180

Mk 1G

R16277

R16367

R16297

WRV 204 110 ->193

Mk 6A

R20737

R20917

R20757

WRVi (H) 225 110 ->193

Mk 6A

VR25210-0

VR25101-0

VR25102-0

WRV 255 220

Mk 5

R251157

R251167

R251177

WRV (T) 255 110 -> 220

Mk 5

R251157

R251187

R151177

WRVi(H) 321 132 -> 193

Mk 6

VR32100-0

VR32101-0

VR32102-0

WRV 321 220

Mk 5

R32857

R32867

R32877

WRV (T) 321 132 -> 220

Mk 5

R32857

R32887

R32877

WRVi365 145- 193

MK1

VR36160-0

VR36211-0

VR36162-0

Please see separate file on this CD-Rom for CAD drawings for all of the above. As part of our commitment to continuous product development, the above information is correct at the time of publication and may be updated in the future.

February 2003

3-49


3.6

WRV Compressor allowable nozzle loadings (Suction) Compressor

Bore

Size

mm

Fxx

Fzz

Fyy

Mxx

Mzz

Myy

163/145

125

2500

2500

6250

1250

1250

3750

163/180

125

2500

2500

6250

1250

1250

3750

204/110

150

3000

3000

7500

1500

1500

4500

204/145

200

4000

4000

10000

2000

2000

6000

204/165

200

4000

4000

10000

2000

2000

6000

204/193

200

4000

4000

10000

2000

2000

6000

255/110

200

4000

4000

10000

2000

2000

6000

255/130

200

4000

4000

10000

2000

2000

6000

255/145

250

5000

5000

12500

2500

2500

7500

255/165

250

5000

5000

12500

2500

2500

7500

255/193

250

5000

5000

12500

2500

2500

7500

255/220

250

5000

5000

12500

2500

2500

7500

321/132

250

5000

5000

12500

2500

2500

7500

321/165

300

6000

6000

15000

3000

3000

9000

321/193

300

6000

6000

15000

3000

3000

9000

321/220

350

7000

7000

17500

3500

3500

10500

365/145

300

6000

6000

15000

3000

3000

9000

365/165

350

7000

7000

17500

3500

3500

10500

365/193

350

7000

7000

17500

3500

3500

10500

510/132

350

7000

7000

17500

3500

3500

10500

510/165

400

8000

8000

20000

4000

4000

12000

510/193

400

8000

8000

20000

4000

4000

12000

February 2003

Forces (N)

3-50

Moments (Nm)


3.6

WRV Compressor allowable nozzle loadings (Discharge) Compressor

Bore

Size

mm

Fxx

Fzz

Fyy

Mxx

Mzz

Myy

163/145

75

1500

1500

3750

750

750

2250

163/180

75

1500

1500

3750

750

750

2250

204/110

100

2000

2000

5000

1000

1000

3000

204/145

125

2500

2500

6250

1250

1250

3750

204/165

125

2500

2500

6250

1250

1250

3750

204/193

125

2500

2500

6250

1250

1250

3750

255/110

150

3000

3000

7500

1500

1500

4500

255/130

150

3000

3000

7500

1500

1500

4500

255/145

200

4000

4000

10000

2000

2000

6000

255/165

200

4000

4000

10000

2000

2000

6000

255/193

200

4000

4000

10000

2000

2000

6000

255/220

200

4000

4000

10000

2000

2000

6000

321/132

200

4000

4000

10000

2000

2000

6000

321/165

250

5000

5000

12500

2500

2500

7500

321/193

250

5000

5000

12500

2500

2500

7500

321/220

300

6000

6000

15000

3000

3000

9000

365/145

250

5000

5000

12500

7500

2500

2500

365/165

300

6000

6000

6000

9000

3000

3000

365/193

300

6000

6000

6000

9000

3000

3000

510/132

250

5000

5000

5000

7500

2500

2500

510/165

300

6000

6000

6000

9000

3000

3000

510/193

300

6000

6000

6000

9000

3000

3000

February 2003

Forces (N)

3-51

Moments (Nm)


3.7

Estimated noise levels

February 2003

3-52


3.7

Estimated noise levels (continued)

February 2003

3-53


3.7


February 2003

3-54


3.7


3600 3000

February 2003

3-55

92 dBA 90 dBA


3.7


TYPICAL MEAN SOUND PRESSURE LEVEL DATA FOR WRVi365 COMPRESSOR AT 1m

To Be Advised

February 2003

3-56


3.7


February 2003

3-57


1 Suction Isolating Valve 2 Suction Non-Return Valve 3 Suction Strainer 4 Screw Compressor 5 Discharge Duct 6 Oil Tank/Separator 7 Discharge Non-Return Valve 8 Discharge Isolating Valve 9 Gas Safety Release Valve 10 Flexible Coupling (Spacer Optional) 11 Coupling Guard 12 Main Drive Motor 13 Baseframe 14 Oil Pump Isolating Valve+ 15 Pipeline Strainer+ 16 Oil Pump+ 17 Oil Pump Coupling+ 18 Oil Pump Motor+ 19 Oil Relief Valve+ 20 Oil Non-Return Valve+ 21 Oil Pressure Control Valve 22 Oil Cooler 23 Oil Filter Isolating Valve (Single Filter) 24 Single Oil Filter++ 25 Three Way Valve 26 Duplex Oil Filters++ 27 Solenoid Valve 28 Oil Level Gauge 29 Oil Heater with Thermostat 30 Suction Temperature Gauge 31 Oil Filter Differential Gauge 32 Oil Temperature Gauge 33 Discharge Temperature Gauge 34 Suction Pressure Gauge 35 Oil/Gas Diff Pressure Gauge 36 Discharge Pressure Gauge 37 Suction Pressure Trip 38 Oil/Gas Diff Pressure Trip 39 Discharge Temperature Trip 40 Discharge Pressure Trip 41 Instrument Board

February 2003

3-58


4

LUBRICATION OIL SYSTEMS

4.1

General The oil and gas exits the compressor and returns to the oil reservoir in the discharge line as 2 phase flow. As the gas/oil mixture enters the combined oil reservoir/separator vessel, most of the oil separates from the gas stream due to the reduction in velocity, the separated oil falling into the reservoir base. In a typical refrigeration system the gas and entrained oil droplets then pass through a stainless steel wire mesh separator element mounted in the upper part of the vessel where further separation occurs, the separated oil falling by gravity into the reservoir sump. The coalescing element is then generally designed so that the gas leaving this vessel contains approximately 50 parts per million by weight of oil. If a lower oil carry-over value is required, eg, in a gas turbine application, then high efficiency secondary coalescing elements can be fitted either within the reservoir/separator vessel or in a secondary vessel. The lubricating oil is mechanically separated from the gas, collected and retained in the oil sump reservoir at compressor discharge pressure. Pumped pressure lubrication is required for both the inlet and outlet end hydro-dynamic radial sleeve bearings and also the slide valve load/un-load cylinder connections. The inlet bearing connection(s) internally supplies oil to the shaft seal, the inlet sleeve bearings and the inlet end thrust balance piston. The outlet bearing connection(s) internally supplies oil to the outlet sleeve bearings, the male and female rotor thrust bearings and the outlet end thrust balance piston. The main oil injection connection supplies the rotor cavity oil for compression cooling. A separately driven gear type oil pump supplies positive oil pressure from a retention oil supply. Manifold oil pressure is maintained at a positive differential pressure above compressor discharge pressure, 30-65 psid (2-4.5 bar), by the higher oil pump flow delivery than compressor consumption, by-passing the excess oil pump flow through a differential pressure control valve, back to the oil reservoir. The differential pressure set point (oil pressure minus discharge gas pressure) is obtained from the computer selection program. Oil pressure differential is maintained constant above discharge pressure under all operating suction and discharge pressures/oil flow consumption. All the oil is cooled to the design oil supply temperature which is maintained constant with a thermal actuated valve. External oil coolers may be water cooled, air cooled, or thermo-syphon. Location of the oil cooler and oil filtration may differ depending on the type of lube-oil system supplied, ie, fully pumped or partially pumped.

February 2003

4-1


4.1

General

4.1.1

Fully Pumped Lubrication System The lubricating oil draws from the sump/reservoir through an oil pump strainer and it is pumped with the excess pump quantity piped to a by-pass regulator to produce the design oil pressure differential above discharge gas pressure measured at the compressor manifold. An oil discharge check valve is applied downstream of the oil pump to maintain the oil supply lines primed in the off cycle. After the oil is externally cooled it then passes through a full flow micronic oil filter and the clean, cooled oil then enters the oil manifold, from which all compressor supply connections receive pressurised oil. The main oil injection consumption rate is self-regulating over all compression ratios. A manual throttling valve is permitted on main oil injection to fine tune the discharge temperature range or decrease injection quantity. Automatic regulating valves are recommended on main oil injection to maintain discharge temperature at a constant set point where discharge dew-point control is required in the oil separator. The main oil injection piping is not required on specific applications, typically heavy gases, providing the discharge temperature is not excessive under all operating conditions. This can be evaluated on the computer selection-rating program. The hydraulic slide valve actuator is also fed with pressurised lubrication. Valves are not permitted to interrupt or jeopardise oil pressure on inlet and outlet bearing connections after the manifold with the compressor in operation.

4.1.2

Partially Pumped Lubrication System It is possible to reduce oil pump and oil filter size requirements with a partially pumped lubrication system. A partially pumped lubrication system is not recommended in applications with a continuous design low differential gas pressure and/or continuous low compression ratios. All the lubricating oil is cooled in the oil cooler, which is supplied at discharge pressure from the oil reservoir. The main oil injection to the rotor cavity is diverted before the oil pump and supplied to the compressor through an individual retention manifold. Main oil injection must be supplied to the compressor at a supply pressure no less than 1 Bar (14.5 psid) below discharge pressure. Since the oil supply pressure is reduced to the main oil injection connection, a modified oil flow option may be required to insure discharge temperature is not excessive. A separate main oil injection check valve is recommended for possible transient start up compressor internal gas pressures. A fine mesh strainer is required after the check valve on the main oil injection connection. The bearing oil which is drawing from oil cooler retention passes through an oil pump strainer and it is pumped with the excess pump quantity piped to a bypass regulator to produce the design oil pressure differential above discharge gas pressure measured at the compressor bearing manifold. Oil pump suction piping practices and filtration should be adhered to as recommended by the oil pump manufacturer. The bearing oil passes through a micronic oil filter and the clean oil then enters the oil manifold, from which compressor bearing connections receive pressurised oil above discharge.

February 2003

4-2


The hydraulic slide valve actuator must also be provided with pumped lubrication. Valves are not permitted to interrupt or jeopardise oil pressure on inlet and outlet bearing connections after the manifold with the compressor in operation. 4.2

Oil Filtration Cleanliness and quality of the lubricant used are of extreme importance in achieving long life and trouble free running. Oil filters with replaceable elements rated at 10-15 microns nominal or 25-30 microns absolute must be used for bearing oil and un-loader actuator supply connections. Dual oil filters with transfer valve is permitted providing oil pressure is not compromised during compressor operation. Filtration requirements for the main oil injection connection (supplies rotor cavity for compression cooling) with partially pumped lubrication systems require oil filtration to 250 microns absolute. Bearing oil and un-loader actuator supply connections require 25-30 microns absolute (maximum particle size). The capacity of the oil filter should be carefully considered to ensure that excessive filter service intervals are not incurred. Factors such as the cleanliness of the gas and oil systems should be taken into account when selecting the filter. A differential pressure indicator and/or pressure switch should be fitted across the oil filter to indicate the degree of contamination of the filter element. A single filter with isolating valves can be used when service interval shut downs are permitted. An oil filter by-pass is not permitted. When continuous operation is required, then duplex filters with a transfer valve should be installed providing oil pressure is not compromised during compressor operation. During construction of the compressor package, it is very important that pipe scale, weld splatter and other debris are removed from the oil and gas systems. During commissioning the oil system should be flushed and proved clean before being connected to the compressor.

February 2003

4-3


4.3

Oil Piping It is essential that small bore piping and/or reduced bore fittings are not used between the compressor and the oil manifold. It is not permitted to restrict the oil flow and/or oil pressure to the compressor bearing connections and the oil manifold. Pressure drop between the oil manifold and compressor must be negligible for balance pistons, sleeve bearings and shaft seal. The inside diameter of the piping must be never less than the connection size on the compressor, eg, where the connection size is ½” BSP, piping with a minimum diameter of ½” must be used. Each bearing supply connection must have a separate supply line from the manifold, ie, no branching is allowed with bearing connections. Main oil injection requires a separate supply line from the manifold, however branching is permitted on casings with 2 main oil connections. When packaging WRVT, WRVTS, WRVTN type compressors oil supply connections to the thrust bearings on BOTH male and female rotor sides must be piped to the oil manifold. Additional information on oil piping for WRVT510 compressors is given in Section 9.

4.3.1

Closed Loop Refrigeration Low temperature applications may require skimmers (jet pump) in the evaporator to return circulating system oil.

February 2003

4-4


4.4

Recommended Lubricating Oils Howden compressors are fitted, as standard, with neoprene ‘O’ rings. These ‘O’ rings are compatible with the majority of standard refrigerants and many oils, but compatibility with all possible combinations of refrigerant and oil cannot be guaranteed. Should you wish to review this matter, please do not hesitate to contact Howden Compressors Limited, who will be pleased to provide recommendations and costs for any special ‘O’ ring materials which may be required. Typical Standard Refrigerants: R717, R22, R134a, R404A, R407C, R410A, R507 Typical Oils: Mineral Oil Polyol Ester Oil Poly Alkylene Glycol Oil Alkyl Benzene Oil Oil Viscosity: For the majority of ammonia refrigeration applications, oil viscosity of 68 centistokes at 40oC is the appropriate selection. However, with many other refrigerants, eg; R134a with high condensing temperatures, or applications involving hydrocarbon gases, a specific oil selection is required. Howden Compressors Limited offers a consultancy service to all users of HCL product. Please consult the applications department of HCL who will be happy to advise on grade of oil applicable to the refrigerant or gas at the specific duty application.

February 2003

4-5


5

SUPERFEED SYSTEMS

5.1

Principle of the Howden Superfeed System The Howden Superfeed compressor is a development of the screw compressor whereby an additional charge of gas is handled by the compressor over and above that which is drawn in through the suction in the normal manner. Principle of Operation The principle of operation is based on the additional gas being passed into the compression space after suction cut-off. The compressor operates in the following manner: 1.

Gas is drawn into the compressor by rotation of the rotors in the normal way.

2.

Once this process (1) is complete the charge of gas is automatically sealed between the rotors and casings by the further rotation of the rotors. This point in the process is called "suction cut-off".

3.

At this point in the compression process a port is provided in the casing which connects into this trapped gas. If gas is available at this port at a pressure above compressor suction pressure then some will flow into the compressor. This gas will be additional to the normal suction charge.

February 2003

5-1


5.2

Application of Superfeed to Refrigeration The vapour compression refrigeration cycle is an ideal application for the Superfeed principle. The standard cycle operates in the following manner (see Fig. 1): 1.

Evaporation of liquid refrigerant to vapour thereby absorbing heat from the media requiring to be cooled.

2.

Compression of the vapour obtained from (1) to a higher pressure or energy level termed the condensing pressure.

3.

Condensing of the vapour back to liquid at this high pressure level, thereby rejecting the heat absorbed during the cycle.

4.

Expansion of the high pressure liquid down to evaporating pressure from where the cycle proceeds as before.

Fig. 1 The Vapour Compression Refrigeration Cycle Represented on a Mollier or Pressure Enthalpy Chart

February 2003

5-2


5.2

Application of Superfeed to Refrigeration (Cont'd) The important part of this cycle to the superfeed principle is (4). It will be seen that the expansion of the high pressure liquid down to evaporating pressure results in a proportion of the liquid vapourising to gas (the "dryness fraction"). If this vapour could be transferred directly from process (4) to process (2) without reduction of the normal compressor throughput, the refrigeration capacity (1) of any given plant would be increased by the quantity "a" as shown in Fig. 2.

Fig. 2 Result of Removal of "Dryness Fraction" on Standard Cycle. The Howden Superfeed Compressor effectively removes the dryness fraction as indicated in Fig. 2 by passing it into the superfeed port in the compressor. Fig. 3 and 4 show the system in a simplified form.

Fig. 3 Simple Superfeed Circuit.

February 2003

5-3


5.2

Application of Superfeed to Refrigeration (Cont'd)

Fig. 4 Simple Superfeed Cycle. However there are a number of practical difficulties in the use of this simple system. A similar result is achieved by subcooling the liquid from the condenser and passing the gas boiled off during this cooling process into the superfeed port, Fig. 5, and this is the system used.

Fig. 5 Subcooling Superfeed Cycle. Summary It can be seen that the Superfeed principle goes a long way towards achieving two stage plant efficiency from a single stage compressor.

February 2003

5-4


5.3

Detail Design Notes on Superfeed Applications 1.

Superfeed Pressure at 100% Capacity

The pressure in the superfeed system is self determining. The compressor can take in a certain volume of gas at any given superfeed pressure. Similarly the refrigeration system develops a certain volume of gas for any given refrigerant mass flow and pressure, and the two balance out at an equilibrium condition. 2.

Superfeed Pressure at Part Load

As the compressor is off loaded by means of the slide valve the superfeed pressure falls towards suction pressure. The overall refrigeration cycle is not affected by this but the system must be designed to ensure satisfactory refrigerant flow to the evaporator which is one of the reasons for using a shell and coil (or equivalent) superfeed vessel. 3.

Detail System Design

The recommended system is that shown in Fig.6. The superfeed vessel is equivalent to a shell and coil intercooler in a two stage system. The shell and coil vessel can be replaced by a shell and tube heat exchanger in some applications as shown in Fig. 6a. A non return valve is necessary in the superfeed line to prevent oil flowing back through this system at shut down. A strainer/filter is also included to prevent any dirt passing directly into the compressor. The cycle with this system is as shown in Fig. 5 and the temperature difference between the subcooled liquid from the condenser and the superfeed vessel liquid should be of the order of 10°F (5.5°C). The temperature of the liquid entering the coil is condensing temperature and the temperature leaving the coil is 10°F (5.5°C) above the "shell" temperature. The actual compressor performance achieved using the "Superfeed" principle is obtained from the Howden computer selection program.

February 2003

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5.3

Detail Design Notes on Superfeed Applications (Cont'd)

Fig. 6 Recommended Superfeed System

Fig. 6a

February 2003

5-6


5.4

Superfeed System Options There are two basic types of Superfeed system, open flash and closed circuit. Both options can be evaluated using the Howden computer selection program and can generally be applied to all refrigerants and most operating conditions. The use of Open Flash is common on Freon 22 as the pressure in the flash vessel is sufficiently high to enable the liquid refrigerant to be expanded down to the evaporator pressure. Using an economically sized expansion valve on Ammonia however, the superfeed vessel pressure is closer to the evaporating pressure and there may not be sufficient pressure difference available. The most common superfeed system for Ammonia is therefore a closed circuit system. For other refrigerants the same criteria apply. Examples of closed circuit Superfeed systems. i) Shell and Coil

Fig. 7 In this system the liquid refrigerant is sub-cooled within a few degrees of the approach temperature of the liquid refrigerant in the shell side. The actual temperature of the subcooled liquid depends upon the design of the heat exchanger.

February 2003

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5.4

Superfeed System Options (Cont'd) ii)

Shell and Tube. (System 1).

Fig. 8

Tubes are immersed in liquid refrigerant which is evaporating at superfeed pressure. The superfeed effect is similar to that of the shell and coil system. iii)

Shell and Tube. (System 2).

Fig. 9

The superfeed effect is again similar to other closed systems, the only difference is that the liquid refrigerant is sub-cooled as it passes over the tubes. Liquid refrigerant is expanded through the tubes and heat is absorbed from the shell side.

February 2003

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5.4

Superfeed System Options (Cont'd) Open Flash Superfeed System.

Fig. 10

Comparative effect of Open Flash and Closed Circuit systems. Where:

h1 =

Enthalpy of vapour at evaporator.

h2 =

Enthalpy of liquid at Superfeed pressure.

h3 =

Enthalpy of liquid at Superfeed pressure (equivalent) plus approach across coil/tubes.

Superfeed refrigeration effect (open flash)

h1 - h2

Superfeed refrigeration effect (closed system)

h1 - h3

Comparative effect: h1 - h3 = factor h1 - h2 The factor gives an indication of the reduction in capacity where a closed system is used instead of an open system. Final selection of the system is made taking account of the performance requirements and the commercial implications of the alternative systems. Full account should be taken of the performance data for each system provided by the computer program. This program incorporates factors which are internal to the compressor and provides a more complete model of performance on which the selection of the most suitable system can be made.

February 2003

5-9


6

COOLING SYSTEMS

6.1

Cooling System Options General Cooling Requirements During normal operation the compressor requires a supply of lubricant to the bearings and seal at an acceptable temperature, typically 40°C (104°F). It is also important, that during compression of the refrigerant, the discharge temperature does not exceed the maximum permitted value (see Section 3.1, Design Limitations). The cooling requirement is generally greater when the compressor is operating at full load. It should be noted that, under certain operating conditions the highest cooling requirement can occur at part load conditions. This can be identified from the computer selection program. The cooling system selected should be capable of removing all the heat added to the oil, either at full load or at part load, whichever is the greater. Oil cooling can be achieved using an external cooler, cooling mediums might include water, refrigerant and air. In many cases the need for an external oil cooler can be avoided by direct injection of liquid refrigerant into the compression space or compressor discharge pipe. This method of cooling can be used provided the compressor discharge temperature and the entrained lubricant can be reduced to a temperature which is below the maximum oil temperature permitted for the bearing oil supply. It must, however, be sufficiently high to prevent condensation of the refrigerant within the oil separator. In both cases the flow of refrigerant is controlled using a thermostatic control such as TEAT valve and solenoid. Liquid Injection into the Compressor A connection for this purpose is provided as standard on all WRV models. There can be a slight reduction in compressor capacity and a marginal increase in absorbed power if direct Liquid Refrigerant Injection (LRI) cooling into the compressor is used.

February 2003

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6.1

Cooling System Options (Cont'd) Liquid Injection into the Discharge Pipe This method of cooling may also be acceptable. In this case the performance of the compressor is not affected, however care must be taken to ensure that all of the liquid injected is fully evaporated before reaching the oil tank/separator. If liquid refrigerant carries over into the tank, high levels of oil carry-over may result. In most cases a longer discharge pipe is required to prevent this problem. Care must also be taken to ensure that the compressor discharge temperature, upstream of the liquid injection point, does not exceed the compressor design limit. When all evaporation has been completed, this method has the advantage that less oil carry-over than normal occurs because the refrigerant temperature discharging into the separator is lower than with other cooling methods.

February 2003

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6.2

External Oil Cooling Oil cooling using an external oil cooler is the most common method used for refrigeration screw compressor applications and the most common cooling mediums are water or liquid refrigerant. Water can either be sourced from an independent supply or if evaporative condensers are used, it may be possible to use water taken from the condenser sump, provided the condenser has sufficient capacity to dissipate the additional heat load. If not, then higher condensing temperatures will result in additional power consumed by the compressor and additional water usage. An advantage of using condenser water is that it is usually treated and will reduce oil cooler maintenance. Whatever the water source, account should be taken of its maximum temperature, the cooler design approach temperature and the maximum allowable oil temperature to the compressor, when the oil cooler is selected. Liquid refrigerant may also be used in an external oil cooler to control the oil temperature. In this case, the liquid is taken from the condenser liquid receiver into the oil cooler and heat from the oil is absorbed by the liquid refrigerant which evaporates. The evaporated vapour may either be fed to the superfeed port of the compressor or in some cases, to the compressor discharge pipe between the compressor and oil separator depending on the operating conditions (consult HCL). In the latter case there must be sufficient head of liquid from the liquid receiver to ensure sufficient thermosyphon effect through the oil cooler. Arrangement of thermo-syphon oil cooling system.

February 2003

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6.3

Cooling Using Liquid Refrigerant Injection Liquid refrigerant is injected into compressors for the purpose of absorbing the heat of compression, thereby eliminating the need for an external oil cooler. Since there is no oil cooler, the temperature of the oil and gas mixture leaving the compressor will also be the temperature of the oil supplied back to the compressor for lubrication. (Radiation and convection heat losses from oil piping and oil separator are ignored since the percentage of heat involved is relatively small). General Limitations There is a maximum temperature at which oil can be supplied to the bearings as dictated by the actual bearing materials, etc. With Howden machines the maximum oil temperature which can be accepted to the bearings is 70oC (158°F). The oil temperature safety cut-out should be set at this temperature and the normal operating temperature set below this to allow for fluctuation during operation. There is also a minimum oil temperature which is set not by materials but by the necessity to ensure that refrigerant does not condense in the oil separator. If the discharge temperature of a compressor, which as noted above is the oil supply temperature, is allowed to fall below condensing temperature then refrigerant will condense in the oil tank/separator. In order to prevent this possibility the compressor discharge temperature must be maintained at least 10°C (18°F) above the condensing temperature. To summarise, therefore, the compressor discharge temperature must stay above condensing temperature but below the bearing maximum oil temperature. Condensing Temperature Limitations with Liquid Injection While the minimum oil temperature can be varied to some extent to suit particular applications, the maximum oil temperature is restrained by the bearing materials and properties of the lubricant. As noted above, the maximum temperature is a trip setting of 70°C (158°F). We would recommend that at least 5°C (9°F) of a margin is left between the normal operating and the trip setting giving a maximum oil inlet temperature of 65°C (149°F). When an allowance is made, therefore, of 10°C (18°F) temperature difference, the maximum condensing temperature at which liquid injection can be used is 55°C (130°F). For all condensing temperatures above this level, liquid refrigerant injection into the compressor is not acceptable.

February 2003

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6.3

Cooling Using Liquid Refrigerant Injection (Cont'd) The Effect on Performance of Liquid Refrigerant Injection Injecting liquid refrigerant into the compressor for cooling purposes can reduce compressor performance. Compared to a compressor using an external oil cooler, the amount of penalty varies with the refrigerant and the operating pressure ratio, and is more severe at higher operating pressures and with Freon type refrigerants. Ideally, the liquid refrigerant should be injected into the discharge mixture leaving the compressor, and would therefore carry out the cooling function without affecting the compressor performance. This, however, would require a liquid refrigerant pump to raise the refrigeration pressure above the compressor discharge/condensing pressure which would add expense to the system. There could also be some problems with oil separation if all the refrigerant is not vapourised by the time it passes through the separator. This method however, can be used with a booster compressor since liquid refrigerant is available at second stage condensing temperature and therefore can be easily injected into the discharge pipe for the first stage. It is important to remember that refrigerant will still condense in the oil tank separator if the temperature after cooling is at or below the condensing temperature. The 10°C (18°F) temperature difference should still be maintained and the 55°C (130°F) maximum condensing temperature limit with this system still applies. Method The quantity of liquid refrigerant injected is controlled by a thermostatic valve sensing compressor discharge temperature which adjusts the refrigerant flow to hold this temperature at the desired level. A solenoid valve is also incorporated into this line to prevent liquid refrigerant flow when the compressor is not operating. Liquid Refrigerant Injection Ports WRV 163 to WRV 365 compressors are equipped with two LRI ports situated on the male rotor side of the compressor. The upper LRI port is defined as the Low Pressure LP port. The lower LRI port is defined as the High Pressure HP port.

February 2003

6-5


7

CONTROL SYSTEM

7.1

Control Philosophy The WRV compressors incorporate an integral sliding valve by which capacity can be automatically and steplessly varied from 100% down to approximately 10%, with related power saving. For refrigeration applications, capacity of the compressor is usually controlled in response to changes in the evaporating temperature or the compressor suction pressure. Control using evaporator temperature is more common where the application involves a single evaporator or chiller. Compressor suction pressure is more commonly used for control purposes where more than one evaporator or chiller is served by one or more compressors. This system is particularly suitable where the system uses a flash drum and a pumped refrigerant circuit. For gas applications, particularly fuel gas feeding a gas turbine, capacity is controlled in response to changes in the compressor discharge pressure. Unlike refrigeration applications where demand changes very slowly, demand in fuel gas applications can change rapidly and suddenly i.e. turbine trip situation. In this situation the slide valve control response is too slow and a secondary gas bypass control is required. This system incorporates a gas pressure control valve and gas bypass piping fitted between compressor discharge and suction. It would operate in parallel with the slide valve control system, taking over control when a quick response is required, with control returning to the slide valve when the turbine running conditions return to normal.

February 2003

7-1


7.2

Two Stage Compression Systems On two stage systems, control of the second stage may be achieved in a similar manner, however the following comments may be relevant: Where there is an intermediate/side load, then it is usual to control the intermediate temperature or pressure at a constant level. If, however, there is no side load, the most efficient compressor arrangement is to vary the intermediate conditions so that both first and second stages are loaded to approximately the same degree. In this way both stages are at their best part load efficiency. This approach is only feasible where the first and second stages are reasonably well matched in terms of relative capacities at the ideal intermediate conditions. Generally the ideal intermediate pressure is found from: Pi =

Pev x Pco Where Pi

= Ideal Intermediate Pressure. Pev = Evaporating Pressure Pco = Condensing Pressure.

The mean condensing pressure may be used as the basis for this calculation rather than the maximum condensing temperature. The reduction in efficiency, due to the actual intermediate pressure varying slightly from the ideal, is generally small compared with the improvement in efficiency achieved when both compressor stages are loaded equally.

February 2003

7-2


7.3

Capacity Control The compressor slide valve is operated by a double acting piston within a hydraulic cylinder which is mounted on the compressor. To move the slide valve, oil from the lubrication system is fed, via loading and unloading solenoid valves, to one side or other of the piston which is connected to the slide valve.

7.3.1

Starting Interlocks The compressors capacity or loading is controlled by a hydraulically operated slide valve. In order to prevent the compressors starting under full load conditions a minimum capacity switch should be connected as an interlock in the motor starting circuit.

7.3.2

Control Lubricating oil is used to move the slide valve on load and off load. Oil is vented to the rotor space when the slide valve moves off-load. The slide valve is held in any intermediate position by locking the oil in the actuating cylinder. The oil supply to and from the actuating cylinder is controlled using a Dead Band proportional controller and four 2 position solenoid valves or two 3 position changeover valves. Referring to Fig.3 on 7.4 To load the compressor, energise SV1 and SV3. This will open SV1 to allow oil pressure to operate the actuator cylinder, and open SV3 in the oil drain line. When the compressor is loaded to 100% SV1 and SV3 may be left energised or they can be de-energised. To partially off-load the compressor SV1 and SV3 are de-energised. SV2 and SV4 are energised to allow oil pressure to off load the actuator cylinder. When the required part load condition is satisfied, SV2 and SV4 are de-energised. This locks the pressure in the actuating cylinder, holding the slide valve in position.

7.3.3

Multimachine Operation When a number of compressors are used within one refrigeration system, it may be necessary to identify when an additional machine is required to meet system demand. This can be inferred from a continuous signal to SV1 of the running compressor. The slide valve can also be controlled manually by the operation of ON LOAD/OFF LOAD push buttons. A MANUAL/AUTO selector switch is normally installed to allow the operator to select whichever mode of control is desired.

February 2003

7-3


February 2003

7-4


7.3.4

Capacity Control

The Linear Position Indicator (LPI) General: An electronic device called a Linear Potentiometer gives an indication of the position of the slide valve which can be used by the compressor control system. The Linear Position Indicator (LPI) is an electronic contact-less displacement sensor inserted into a sensor well which allows the LPI to be removed from compressor without loss of oil or gas from the compressor. The LPI has several usable options built into one device. The slide valve position can be indicated in three different ways: Visual Light Emitting Diodes (LED) Visual and by an analogue output 4-20mA Visual and by a digital 24 V DC signal output on minimum and maximum slide valve position

Part Load LED

Max. Load LED

Min. Load LED

Calibration Button Cover

Visual: It is always possible to see the position of the slide valve. At minimum load a green LED is illuminated at the lowest light on the left of the LPI. At maximum load all red LED are illuminated. At part load only some of the LED are illuminated, eg, at 50% load only half of the red LED will be illuminated. NOTE: The compressor can only be allowed to start with the slide valve in the minimum load position. Therefore a signal from the minimum load electronic position switch is always required or if the 4 – 20 mA signal is being used, then a 4 mA signal is required.

February 2003

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7.3.4

Capacity Control (cont’d)

Visual and by an Analogue Output (4-20mA): The minimum load position is given by the 4 mA output and the maximum load position is given by the 20 mA output. (White Wire) Part load positions are indicated by intermediate values between 4 and 20 mA. The LED’s on the indicator also give a visual indication of part load operation. It should be noted that part load slide valve position is not a direct indication of actual compressor capacity at part load. Use of the 4 – 20 mA signal is common for many control systems and may be used on its own, if required, for all control functions for single and multiple compressor installations, subject to a suitable control system. Connections: Wiring Plug Connections 1=Brown 2=White 3=Blue

Function Supply Voltage + 24V DC Output Signal 4-20 Ma Common – 0 VDC

Visual and by a Digital 24V Output on Minimum and Maximum load: There is also another option that can be used to control and get the minimum signal for start-up. This option works the same as the mechanical micro-switches but instead uses the electronic switches incorporated in the LPI unit. These electronic switches give a 24 V DC output. A digital output is given on the Minimum and Maximum position of the slide valve and an interposing relay, which must be incorporated in the control panel in place of each mechanical micro-switch, is activated by the digital signal completing the control circuit signal. This interposing relay must have contacts with suitable ratings. The interposing relay replaces the original switch function. The LED’s only give a visual indication of the slide valve position. If the slide valve is in the minimum position and the LED for minimum is illuminated, there will be a digital output on the green/yellow wire. If the slide valve is in the maximum position and all the LED’s are illuminated, there will be a digital output on the black wire. Existing installations equipped with the mechanical micro-switches can use this option. Connections: Wiring Plug Connections 1=Brown 3=Blue 4=Black 5=Green/Yellow

Function Supply Voltage + 24V DC Common – 0 VDC Digital Output Max. Load Digital Output Min. Load

Choose the best way for giving a start signal and connect the wires according to the table.

February 2003

7-6


7.3.4


Checking the LPI Calibration All compressors with variable Vi are despatched from Howden with Vi set at 2.2 and the LPI calibrated to suit Vi 2.2. When the Slide Valve is in the unloaded position, the 10% minimum load LED should be illuminated. To check that the LPI indicates maximum load when the slide valve is in the fully loaded position, ie all LED’s are illuminated, the following checks should be made. Move the slide valve to 100% (by using the oil pump or if the system is shut down, use a manual oil pump or air pressure). By pressurising the outboard side of the actuator piston the slide valve will be moved to the fully loaded position. The LPI should indicate 100% by illuminating all LED’s. If this is not the case please do the calibration again as follows: Move the piston to the minimum load position. Adjust the Vi screw to set VI to suit operating conditions. Note: This adjustment must only be done when slide valve is in the minimum load (10% position). Remove the calibration button cover, Switch power on and disconnect the electrical plug under the LPI. Wait for 2 minutes. Re-connect the plug. All the capacity array lights and red LED at the calibration button will light for 2/3 seconds and then go out. After approximately 20 seconds, the green LED light at the calibration button will start flashing (the minimum capacity array light may come on). Allow 5 minutes to elapse before starting calibration. To start calibration, press the calibration button once. The green LED by the calibration button will go off and the red Led by the calibration button will come on steady for approximately 15 seconds and then start flashing. Now move the slide valve to the maximum load position. During this movement the cylinder capacity array lights will start to light. When the slide valve is in the max. load position, push the calibration button once. The red LED by the calibration button will stay on for approximately 15 seconds and then go off. The green LED light will come on, possibly flashing. The calibration is now complete. Refit the calibration button cover. Note: If the operating Vi is changed, then the LPI will have to be re-calibrated as from 6 above.

February 2003

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7.3.4


The LPI Linear Position Indicator

Calibration Button Green LED

Red LED

LPI Wiring Diagram

February 2003

7-8


8

COMPRESSOR AUXILIARIES

8.1

Compressor Drive Coupling A flexible coupling must be used which may be either of the rubber block type or the metal membrane type. Coupling alignment is extremely important and must be within the manufacturer's alignment tolerances. It must be checked before any pipes have been fitted to the compressor and rechecked after all pipes have been fitted. If a change in alignment occurs after the pipes have been fitted, the compressor is being strained. Pipes and supports must be adjusted to prevent this from happening and the compressor must not be run until alignment readings taken before and after all pipes have been fitted are identical. A spacer coupling may be used where access to the compressor shaft seal is required without moving either the compressor or the main drive motor. In the event that a torsional analysis of a complete compressor and driver system suggests that a change in natural frequency would be advantageous, the torsional stiffness of the coupling would be the area to be changed.

February 2003

8-1


8.2

Gas Suction Strainer A gas suction strainer should be included in the system to capture any particles of dirt, corrosion, weld spatter or anything of damaging size from the gas before it enters the compressor. Proprietary brands of pipe line strainers are available from specialist manufacturers such as Herl and Henry and should be designed so that when the strainer is removed for cleaning, all the dirt etc is contained inside the strainer basket and is removed with it, i.e. gas flow from inside to outside. Many compressor packagers manufacture their own gas suction strainers and the following guidelines are given. The gas strainer should have gaps no greater than 0.25 mm (0.010"). A suitable wire mesh is 37 S.W.G. (0.173 mm dia., 0.0068" dia) and 24 wires/cm (60 wires/inch) giving a gap size of 0.24 mm (0.0096"). This fine mesh is too light in construction on its own and requires support from heavier material. One method of support is to sandwich the fine mesh between two sheets of heavier mesh of 20 S.W.G. (0.91 mm dia., 0.036" dia) and 8 wires/cm (20 wires/inch). This "sandwich" can be formed into a plain or corrugated cylinder as shown in the sketches below. It is important that the free area of the fine mesh should be not less than 2.5 times the area of the incoming suction pipe. The perforated plate support should have a free area not less than 1.5 times the pipe area. A typical construction of suction strainer is shown below. This allows the strainer basket to be removed for cleaning without disturbing either the pipework or compressor.

February 2003

8-2


8.3

Compressor Suction and Discharge Non-Return (Check) Valves These valves are fitted to prevent the compressor rotating back due to the high pressure gas on the discharge side when the plant is stopped. The suction check valve also prevents oil flowing back down the suction line. This can happen if either the non-return valve is fitted to the compressor discharge only and the oil pump is operated, or when the compressor is shut down and the oil in the oil cooler etc. degasses and pushes oil into the compressor. The selection of suction non-return valve type is important as pressure loss at this point in the system has a significant effect on the overall performance. Disc type non-return valves ensure low pressure losses. The split disc type is also suitable. It has been found that some disc types where the disc is mounted on a spindle, can suffer from chatter at part load unless some damping is provided. An 'O' ring fitted to the spindle can solve this problem. The suction non-return valve should always be positioned before the suction strainer. This is to ensure that should any part of the valve come loose it is caught in the strainer and does not pass into the compressor. The size of the non-return valve should be the nominal size of the suction ductwork. The maximum recommended gas velocity through the suction system is 15 metres/sec. (50 ft/sec). When handling very light gases such as helium (He) pressure losses are low and higher velocities are acceptable.

February 2003

8-3


8.4

Discharge Pipe Connecting Compressor to Oil Tank/Separator The diameter of this pipe is sized to suit the duty involved, i.e. the velocity flowing through it. The maximum recommended velocity is 25 metres/sec (80 ft/sec). The pipe size will, in many cases, be less than the bore of the compressor discharge port. This is because the compressor port is sized for the highest velocity duty possible within its design specification and most duties are lower than this maximum. To achieve the lowest losses between the compressor discharge port and the discharge pipework, a tapered pipe reducer is preferred to a plain flange. Instrumentation (and safety cut outs) measuring compressor discharge pressure and temperature are normally connected into this pipe section.

February 2003

8-4


8.5

Oil Tank/Separator This is a pressure vessel and must be designed to the requisite code. It may be designed as either a vertical or horizontal unit. In the case of the vertical design, the separator is a separate item usually mounted on the baseplate which also carries the compressor and its drive unit, the oil pump(s), oil cooler, oil filter(s), all the interconnecting pipework and sometimes the instrument/control panel. In the case of the horizontal design, the separator usually forms part of the structure and has the compressor and driver mounted on it with the other auxiliaries either attached or located around the external structure. A frame round the vessel transfers much of the weight of the major components to the supporting feet. Separator Vessel Diameter - Oil Reservoir Capacity This vessel has two main functions. It contains the oil reservoir and separates the oil from the gas before it passes to the rest of the system. The diameter of the vessel has to take both of these functions into account and is normally set by the area of the demister and the height, as required for the necessary oil volume and other physical considerations. The first stage in the oil separation process is to arrange for oil which is in large droplet form to return by gravity to the main oil reservoir in the bottom of the separator vessel. At least 97% of the oil is separated out in this way. The method of feeding the gas/oil mixture from the compressor into the separator is carefully designed to reduce velocity, reverse flow direction etc. to achieve maximum effect from this first separation stage. Details of how to do this are provided later in these notes. The quantity of oil in the reservoir should be a minimum of 114 litres (25 Imperial gallons) or half the oil pump capacity per minute whichever is the greater. Separator Element Selection There are two main alternative types of separator concepts in use, and sometimes both are used in combination when extra high efficiency of separation is required. The two types are wire mesh pads and coalescing cartridges (sometimes called cylinders or candles).

February 2003

8-5


8.5

Oil Tank/Separator (Cont'd) Wire mesh pad separators are self cleaning in operation, have very low pressure drop, and require no maintenance or replacement. Separator vessels therefore, do not need flanged ends and achieve separation levels down to 30 to 50 ppm by weight. Coalescing separators have a higher pressure drop (typically 0.3 bar or 4 psi), and do require to be replaced after one to two years in service. Separators must therefore have flanged ends for access to change the elements. By using both systems in combination, separation levels down to 2 to 5 ppm can be achieved. The following design comments apply to vertical separator design. The knitted wire pad operates in the following way. As the gas/oil mixture flows through the pad, the small oil droplets impinge or catch on the strands of wire and build up to large droplets as more small drops collect. Once large enough they fall off the bottom of the pack into the oil reservoir. The gas with the oil removed carries on through the top of the pack. It is important, therefore, to have a thick enough pad to ensure that there is time for all the small droplets to come in contact with the wire mesh and also that there are as many wires as possible in the pad to achieve maximum collection of droplets. It is also necessary to ensure that the ideal velocity of the gas through the pad does not exceed the separating effect of gravity on the oil in the gas. If this does occur the pad does not drain and very high oil carry over results. The knitted wire pad should be made from 2 ply stainless steel wire of 38 S.W.G. (0.152 mm dia., 0.006" dia) the pack having a density of 0.192 gm/cm³ (12 lb/ft³). The pack thickness should be not less than 300 mm (12"). The pack should be produced in the form of a continuous knitted sleeve which is half crimped and then rolled compactly into a circular pack. This would be fitted with grids above and below which provide basic support for the pack in operation. The grids would be smaller in diameter than the vessel but the pack itself should be a tight fit in the vessel to ensure that all gas/oil mixture has to pass through the pack and cannot by-pass it. There should be no openings through the pad, either because the pad is not rolled tightly enough or because the attachment of the grids to the pack (usually these are tied to the pack by wire) distorts it to produce a through opening.

February 2003

8-6


8.5

Oil Tank/Separator (Cont'd) Calculation of Wire Mesh Pad Velocity The area of pad required is calculated by dividing the volume flow of gas passing through the separator by the design velocity for the given conditions. Pad area required = Volume Flow = Velocity

m³/sec = m² m/sec

OR ft³/sec = ft² ft/sec

The volume flow passing through the separator is the compressor mass flow multiplied by the specific volume of the gas at discharge pressure and temperature. The design velocity for the wire mesh pad is calculated from the following equations. Metric Units Design vapour velocity, metres/sec = 0.7 x 0.11

Pi - Pv Pv

Where PL is density of entrained liquid Kg/m³ PV is density of vapour Kg/m³ Imperial Units Design vapour velocity, ft/sec = 0.7 x 0.35

Pi - Pv Pv

Where PL is density of entrained liquid lb/ft³ PV is density of vapour lb/ft³ The design velocity calculated by the above equation incorporates a safety margin to cope with variation in velocity which can occur during normal operation.

February 2003

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8.5

Oil Tank/Separator (Cont'd) These knitted pads can handle a range of velocities below the design value calculated by the equation on page 8.5.03. This makes them suitable to handle the varying volume flow from the Howden variable capacity compressor and permits a number of standard sizes of separator to be designed, each of which can handle a range of volume flows. To assist with this concept of standard components, and to permit the oil volume contained to match the system requirement without the vessel being excessively tall, a number of standard pack areas can be designed to fit into each vessel. The simplest way to reduce the effective area of the wire pack is to mask part of it. The entry of the gas/oil mixture into the vessel is important in order to achieve effective preliminary separation and also in relation to the wire pack position and shape. A simple and effective method is shown below.

This design causes the gas/oil mixture firstly to flow down towards the main oil reservoir and then to reverse direction upwards at low velocity through the wire pack and out of the separator vessel. The large oil droplets, which comprise nearly all the oil, do not reverse direction and tend to continue to flow down in to the oil reservoir, assisted by gravity.

February 2003

8-8


8.5

Oil Tank/Separator (Cont'd) The cover or pipe cap is held in place by webs welded to the pipe end. In order to avoid unnecessary pressure losses the annular area for the gas flow B should be not less than 1.5 times the pipe area A.

In this case the Gas Flow Area B and the Gas Flow Area C should be not less than 1.5 times the pipe area A. It is important that the pipe cover is not positioned too close to the wire pack or it will mask part of it and cause excessive velocity through the remainder. The following rules should be applied.

The distance L should be 0.5D or 350 mm (14") whichever is the greater.

February 2003

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8.5

Oil Tank/Separator (Cont'd) Another arrangement can be used whenever the full area of the vessel is not needed for the wire pack and has the advantage that the overall vessel height is reduced.

In all of these arrangements, the gas/oil mixture first flows towards the oil surface and then reverses direction. If the velocity against the oil surface is too high, there is the danger that re-entrainment of oil can occur causing more oil to pass through the separator instead of less.

The height 'h' should be not less than twice the diameter of the incoming pipe.

February 2003

8-10


8.5

Oil Tank/Separator (Cont'd) The position of the discharge pipe leaving the separator vessel is also important. If this is located too close to the separator pad or at the side there will be preferential flow towards it and much of the pad area can become ineffective, resulting in high oil carry over. It is suggested that the layout illustrated in the following sketch be used:

The length 'h' from the top of the separator pad to the top of the vessel should be 0.5D or 450 mm (18") whichever is the greater. The space 'S' should not be less than the diameter of the exit pipe i.e. not less than 'd'.

February 2003

8-11


8.5

Oil Tank/Separator (Cont'd) Oil Level in Oil Tank/Separator The oil level in the oil tank separator must be arranged so that the compressor rotors are above the oil level at start up. If the rotors are under the oil level during starting this causes a much increased starting torque. The best arrangement is for the oil level to be slightly below the compressor discharge flange. This not only ensures that the rotors are clear of the oil but allows the compressor to be removed from its position without having to drain oil from the oil tank/separator first.

Some form of oil level indicator should be fitted to the oil tank separator which can be a sight glass or "level eyes" i.e. small transparent windows. When the refrigerant is Ammonia, which is less dense than oil, it can condense in the sight glass and sit on top of the oil in the glass tube. If the top of the ammonia level is used as a guide, it will be reasonably accurate but will show a level slightly higher than the oil level internally.

February 2003

8-12


8.5

Oil Tank/Separator (Cont'd) Pipe from Compressor Discharge into Oil Separator As shown in earlier sketches the discharge pipe from the compressor usually enters the vessel and bends upwards to the point of discharge within the separator. In operation this pipe can contain quite a lot of oil giving a misleading indication of oil level in the oil tank. When the compressor is at part capacity much of the discharge pipe can be filled with oil. At full or close to full capacity this oil is blown through into the oil tank separator. In order to minimise these effects an opening in the pipe as indicated should be provided. This opening does not cause any problems with oil.

February 2003

8-13


8.5

Oil Tank/Separator (Cont'd) Oil Pump Suction in Oil Tank/Separator In order to allow the oil tank/separator to be mounted as low as possible it is normal for the oil pump suction pipe to be located as indicated:

The oil pump suction pipe should not be located in the centre of the vessel. The centre is the lowest point in the tank and is the area in which any dirt etc. will tend to gather. It is better to have the pipe offset a little to avoid drawing in any dirt. Piping systems for oil supplies to the compressor should be designed to limit pressure losses to < 1 bar and reduce foaming.

February 2003

8-14


8.5

Oil Tank/Separator (Cont'd) Oil Filling - First Charge Before any compressor set can be started, the oil tank has to be filled. There are two ways of doing this. A connection can be provided fitted with a blank flange, as shown below, which allows either gravity or pumped filling. Alternatively one of the other connections such as the oil relief valve return can be disconnected and the oil fed into the tank via this connection by a pump through a flexible pipe. A separate oil filler connection is preferred. Oil Addition in Operation It is necessary to be able to add oil to the system while it is pressurised with gas or refrigerant. This can be done in three ways: 1.

Topping up Tank

A small tank of capacity of not less than 5 litres (1 Imperial gallon), or 10 to 15 litres (2 to 3 gallons) in larger sets is mounted on the side of the oil tank/separator. This small vessel, which is designed to be suitable for the system maximum pressure, has connections top and bottom, through isolating valves, into the oil tank/separator. It also has a connection in the top, flanged or screwed, of, say 50 mm (2") diameter. For convenience of operation this top connection can be fitted with a valve rather than a blank flange. In order to add oil to the system, the valves on the two pipes into the oil tank/separator are closed and the topping up tank is filled through the large connection on the top. The top connection is then closed, the two side valves opened and the oil flows into the main tank by gravity. This process is repeated until the required quantity of oil has been added.

February 2003

8-15


8.5

Oil Tank/Separator (Cont'd) 2.

Pumping in Additional Oil

Additional oil can be added by pumping it into the system. It is obviously easier to pump into the low pressure side of the system which is the compressor suction. A fairly small connection fitted with a valve is mounted on the suction strainer. A flexible pipe from an oil pump is connected to this and oil can then be pumped into the system. As the oil is being added into the compressor, not directly into the oil tank, it is convenient if the compressor is running as this will pass the oil quickly to the oil tank allowing a check to be made of the correct quantities by means of the oil level indicator. A hand or motor driven pump capable of developing the required pressure may be used. Only clean oil should be added in this way as it is passing directly into the compressor. 3.

Drawing in Oil By Low Suction

An even simpler method of adding oil is to draw in the oil directly. If the compressor suction pressure is below atmospheric pressure the same valved connection described above is used. A flexible pipe is connected and the other end placed in a container of CLEAN oil. When the valve is opened oil will be drawn into the system. When the compressor suction pressure is above atmospheric pressure, the compressor suction isolating valve has to be closed sufficiently to reduce it to below atmospheric pressure to approximately 0.5 atmospheres absolute. It must not be closed completely or cavitation in the compressor will occur. Once this sub-atmospheric pressure is reached, the small valve can be opened to draw in oil as described above. CAUTIONARY NOTE: One of the disadvantages of this system is that it is very easy for the flexible pipe to come out of the oil or for the oil container to empty and air is then drawn into the system. If not noticed, quite large amounts can be drawn in and will have to be purged out of the condenser in a refrigeration system. Care has to be taken that the main suction isolating valve is fully opened when the addition of oil is complete.

February 2003

8-16


8.5

Oil Tank/Separator (Cont'd) Oil Drain from Oil Tank/Separator A drain connection should be provided at the lowest point in the oil tank, which is at the centre of the domed end. This should not be too small so that it can be used for flushing out the vessel prior to operation and for draining out the oil whenever it requires to be replaced. It is recommended that the drain should be not less than 25 mm (1" dia) and should be larger on large capacity sets. In a simple set the drain may be just a plug, but it is often convenient to pipe the drain to a valve for ease of use. Where height is restricted and it is necessary to mount the vessel as low as possible, an internal drain pipe may be used, although it must be remembered that with this arrangement full draining will only be achieved using a pump. Where an internal pipe drain is used the end of the pipe should be close to the lowest point, and the recommended level is 25 mm (1") above the tank base.

Simple drain plug

Drain piped to isolating valve

Internal drain pipe

When the internal drain pipe is used it is recommended that a drain plug is still fitted to allow proper flushing of the vessel prior to its installation on the baseplate.

February 2003

8-17


8.6

Coalescing Oil Separators It is sometimes necessary to reduce oil carry over to lower levels than can be achieved with wire mesh designs alone. This is achieved by using coalescing separators. These are very fine materials through which the gas/oil is passed and the very fine oil mist is "coalesced" into large droplets which are collected downstream of the element. These elements are very efficient but have a high pressure drop, typically 0.3 bar (4 psi) and require to be replaced occasionally. These coalescing elements can act as secondary separators after the wire mesh pad but mounted in the same vessel, in a secondary vessel or sometimes instead of the wire mesh pad in the primary vessel. These three options are shown diagrammatically below.

Information concerning the number and size of the coalescing cartridges (sometimes called elements or candles) is provided by their manufacturer. In arrangement (1) the space 's' between the knitted wire pad and the plate holding the candles should be not less than 300 mm (12"). In arrangement (3) the space 'x' between the deflector and the plate holding the candles should be 0.5D or 350 mm (14") whichever is the greater.

February 2003

8-18


8.6

Coalescing Oil Separators (Cont'd) Oil Return from Coalescing Separators It must be remembered that oil separation by coalescing cartridges collects downstream of the elements, i.e. after the elements. Because there is a pressure drop through the element, typically 0.3 bar (4 psi), this oil cannot simply be drained back to the oil tank. It has to be returned to the compressor. A connection is provided on all Howden compressors which returns this oil to the rotors after cut-off so the compressor performance is not affected. (See Arrangement Drawings in Section 3). There are two ways of returning this oil back to the compressor. The first relies on a continuous but restricted flow back to the compressor, and the second uses a small vessel with level control.

February 2003

8-19


8.6

Coalescing Oil Separators (Cont'd) System 1 is very simple but a balance has to be made between a very small orifice diameter controlling the flow and a small orifice diameter being easily flooded. It is for this latter reason that a filter is fitted upstream of the orifice. A typical orifice diameter for 163 or 204 size compressors is 3 mm (1/8"). Sometimes it is possible to use small bore piping i.e. 5 mm or 6 mm bore (1/4" bore) without an orifice as the restriction of the pipework is sufficient to return the oil without excessive gas returning also. System 2 is a more positive system and should be used for all large compressor sets. With this system oil from the coalescer section is drained out into a small pressure vessel. This vessel is fitted with a level controller which operates a solenoid valve to return oil from this vessel to the compressor. Discharge of Gas/Refrigerant from Coalescing Separators The discharge pipe from the coalescing section has to be carefully positioned to avoid preferential flow causing high local velocities. If the pipe is simply positioned out of the side of the vessel adjacent to the candles or elements there will be an automatic pattern of flow towards that point. This results in gas flowing preferentially to the elements closest to the discharge pipe. The gas velocity through these is higher than elsewhere and in fact the elements furthest away can have very little flow passing through them. The result of the high velocity through some elements is to reduce their separation efficiency and can also cause some of the oil separated from the gas to be re-entrained. A baffle should be positioned over the outlet pipe to prevent this problem arising.

February 2003

8-20


8.6

Coalescing Oil Separators (Cont'd)

An alternative and simpler approach is shown in view (2). This gives an ideal gas flow pattern but does involve a slightly taller vessel and it must be designed to permit the coalescer elements to be replaced. These coalescer elements are not self cleaning and over a period of time the pressure drop across the elements increases and they require to be replaced. The separator vessel design, therefore, must be flanged to permit access to the elements for replacement. Separator Discharge Non-Return (Check Valve) A non-return (check) valve is always fitted in the gas line after the separator. This valve ensures that the compressor pressure is at low level for starting.

February 2003

8-21


8.7

Oil Heater It is often necessary to fit an oil heater into the oil tank for one or both of the following reasons. 1.

To prevent accumulation of liquid refrigerant in the oil tank during shutdown, particularly in the case of freon refrigerants (R22 etc).

2.

To prevent attempts being made to start and operate the compressor with oil so cold and viscous that it will not circulate properly.

The heater should be located below minimum oil level. It should be fitted with its own temperature sensing and control. In order to avoid overheating the oil at the heater surface, the maximum heat rating should be 1.25 watt/cm² (8 watt/in²). The minimum temperature for the oil tank should be 35°C. The heater rating has to be calculated on the basis of the time required for heating the contents of the oil tank and the capacity of the tank. If the time is not specified for any given installation it is suggested that a time of approximately 4 hours is used. Sometimes more than one heater has to be fitted to comply with the combination of maximum heater surface rating and time to raise oil tank to temperature.

February 2003

8-22


8.8

Pressure Relief Valve As shown on the system diagram a non-return (check) valve is fitted at the discharge of the oil tank separator. Most pressure vessel regulations require that a relief valve be fitted whenever a vessel is located between a positive displacement compressor (which includes a screw compressor) and an isolating valve. A relief valve should therefore be fitted to the oil tank separator, capable of passing the full compressor capacity. It is normal for the relief valve to be located after the separator element.

This relief valve can pass gas to the compressor suction (upstream of the suction strainer) or to some other part of the system. This is not the main system relief valve in a refrigeration plant which must be located elsewhere in the system.

February 2003

8-23


8.9

Compressor By-Pass There can be two reasons for incorporating a by-pass line from the oil separator back to the compressor suction. The first application is where it is necessary for the compressor to be off-loaded completely, that is to zero flow (0%). The slide valve incorporated into the compressor reduces flow to a swept volume of 10% which gives an actual volume somewhat less than this but not normally to 0%. The reason why the swept volume is not reduced to 0% is that it is necessary to have a positive swept volume to ensure that the oil being fed to the compressor is returned to the oil tank at all times. When the slide valve is fully opened to its 10% position, the solenoid valve in the by-pass line is opened reducing flow to the system to zero. This valve need only be sized to handle 10% of the mass flow and is therefore quite small in size. The other reason for fitting a by-pass valve is for applications which require the compressor capacity to be reduced almost instantaneously but where it is desirable for the compressor to be kept running. A by-pass line sized for the full mass flow will achieve this and the normal off loading by means of the slide valve will occur in the normal way over a slightly longer period. Although the by-pass valve has to pass the full mass flow, a high pressure drop is acceptable and therefore the by-pass line can be much smaller than the compressor discharge line. It is worth noting that a by-pass line can ensure the minimum possible starting torque. The compressor should always have the slide valve in the fully off load position for starting and starting torque is therefore low. With a by-pass line there can be no build up of pressure during the starting phase. However this is rarely needed.

February 2003

8-24


8.10

Oil Pump Suction The location of the oil pump suction pipe within the oil tank/separator has already been discussed. In the most economic compressor set the oil pump suction will be connected directly to this pipe. However, it is possible that weld scale or corrosion may exist within this pipe and come loose, or that some particles or dirt be drawn in from the oil tank which might damage the oil pump. An oil pump suction strainer is therefore fitted to protect it. It should be of the wire basket type intended to remove large particles only, the fine filtration taking place later in the system. The mesh in the strainer should have a gap size of the order of 500 microns (0.020 inches). It is important that too fine a mesh is not used as this can cause oil pump cavitation particularly with refrigerants that are highly soluble in oil, such as Freons. If it is necessary to be able to clean the oil strainer without depressurising the whole system, isolating valves have to be fitted. One would be located before the strainer and the second after the oil pump. A pump relief should be fitted at the pump discharge, upstream of the isolating valve. This arrangement permits the oil pump to be maintained without depressurising the total system.

February 2003

8-25


8.11

Oil Pump A positive displacement oil pump, such as a gear pump, should be used according to recommendations in Section 4. A centrifugal oil pump should not be used. The capacity of the oil pump is obtained from the computer programme. A full flow relief valve must be fitted at the pump discharge, upstream of any isolating valves.

February 2003

8-26


8.12

Oil Cooler The purpose of the oil cooler is to reduce the temperature of the oil before it is supplied to the compressor for lubrication and cooling of the gas being compressed. If liquid refrigerant injection is being used either into the compressor or the discharge pipe a conventional oil cooler is not needed. There are three principle ways of cooling the oil using a conventional shell and tube oil cooler. The first uses straightforward water cooling. The second uses refrigerant cooling with a thermosyphon arrangement for circulation. With this system the lowest connection on the cooler is connected to the bottom of the refrigerant liquid receiver and the upper connection to a point high up in the liquid receiver. It is necessary for the liquid receiver to be located at a higher level than the oil cooler and the sizing of the refrigerant pipes is important. Also, the cooler and receiver must not be too far apart. If a small refrigerant pump is fitted in the circuit, flow is guaranteed and the system design is much simpler. The third system uses a refrigerant cooled oil cooler but with the refrigerant flow being controlled by an expansion valve varying flow to keep the oil temperature leaving the cooler at a controlled value. The refrigerant leaving the cooler is at a low pressure and may usually be passed to the compressor superfeed connection if superfeed is not being used. In the vast majority of applications a straightforward water cooled oil cooler is used. In some cases cooling water is either in short supply or simply not available. In a refrigeration plant refrigerant oil cooling can be used as mentioned above. Another alternative is air cooled oil cooling. This involves a heat exchanger with air being blown across it to remove the heat. This is usually quite practical but air cooled heat exchangers are usually more expensive than water cooled designs. If an air cooled system is sometimes required to operate with very low ambient temperatures, it is recommended that a thermostatic by-pass is fitted. This is set to control the oil temperature to the desired level and will by pass the cooler until the oil heats up.

February 2003

8-27


8.13

Oil Filter Oil filters of the replaceable element micronic type are necessary with a nominal rating of 10-15 microns (nominal). The simplest system uses a single oil filter with isolating valves on either side of it. This, however, requires the compressor to be shut down whenever the filter element has to be replaced. With a duplex arrangement the compressor can continue to operate on one filter whenever the second is being changed. Four isolating valves may be used or an automatic change over valve.

It is recommended that a differential oil pressure gauge is fitted across the oil filter(s) to give an indication of when the elements need replaced. If a differential gauge is not available a pressure gauge before and after the filter will serve the same function.

February 2003

8-28


8.14

Oil Manifold It is essential that the oil manifold as well as all oil piping downstream of the oil filter is thoroughly cleaned from dirt, dust and weld spatter, as contamination is liable to be loosened by the refrigerant which will then pass directly into the compressor. The oil manifold should have a blank flange at the opposite end from the oil supply connection to enable through cleaning to be carried out. The diameter of the manifold should be larger than that of the oil supply pipe to ensure full oil supply to all connections coming from it. The connections will comprise all those required by the compressor, including its hydraulic capacity control system plus all pressure and temperature connections. It should be noted that the piping from the manifold to the compressor should have a bore no smaller than the nominal bore of the compressor connection, e.g. if a compressor connection is 1/2" BSP then the pipe bore must be at least 1/2" diameter. On the small diameter pipes from the manifold to the compressor, screwed connections are used and because some gases and refrigerants are very difficult to seal, 'O' ring connections are recommended. In some specifications all oil piping downstream of the oil filter must be manufactured from stainless steel. The comments on cleanliness etc. still apply to this material.

February 2003

8-29


8.15

Instrumentation Where the compressors are being used with standard refrigerants, refrigerant pressure gauges should be used on compressor suction and discharge areas, i.e. gauges with dual pressure and corresponding saturation temperature scales. All pressure gauges in a refrigeration system must be able to withstand the maximum static pressure that the system can develop plus a suitable operating margin. The following gauges should be fitted as minimum: Compressor Suction Pressure Oil Manifold Temperature Compressor Discharge Pressure Discharge Pressure (downstream of separator outlet) Oil/gas differential pressure gauge Oil pressure before filter The high pressure side of the differential pressure gauge should be connected to the oil manifold and the low pressure side to the compressor discharge. If it is not possible to obtain a suitable differential pressure gauge this can be replaced by a normal pressure gauge in the oil manifold. The differential pressure of the oil relative to the compressor discharge pressure can then be obtained by the difference between the two gauge readings. It is also possible to replace the oil pressure gauge before the oil filter with a differential pressure gauge across the oil filter giving a direct reading of oil filter pressure drop. Care should be taken to ensure that instrument sensing points are not placed in dead ends of piping.

February 2003

8-30


8.16

Safety Trips The following are the minimum requirements: Compressor Discharge Pressure - Tripping on rising pressure Compressor Discharge Temperature - Tripping on rising temperature Oil Manifold/Compressor Suction Pressure - Tripping on falling differential pressure As mentioned elsewhere in these notes, an oil pressure relief valve and a system gas pressure relief valve are required.

February 2003

8-31


8.17

Instrument Piping All instrument and safety trip small bore piping should be stainless steel. The only exception is for freon refrigerant applications where either stainless steel or copper may be used.

February 2003

8-32


8.18

Driver Requirements In the majority of applications the compressor driver is an electric motor. The static torque and run-up torque required by the compressor are given in Section 3.4 of this manual. The acceleration will depend upon the torque available from the driver and the starting method, in the case of an electric motor driver. Other suitable drivers include reciprocating engines, gas and expansion turbines. In all cases, the starting torque can be minimised by the use of a gas bypass as described in Section 8.9.

February 2003

8-33


9

SPECIAL INSTRUCTIONS FOR WRVT510 COMPRESSOR PACKAGING 1)

Howden Compressors strongly recommend that torsional and lateral analyses are carried out on any system incorporating a WRVT510 compressor.

2)

Howden Compressors strongly recommend that dynamically balanced, metal membrane type drive couplings are utilised in WRVT510 compressor packages.

3)

If the drive motor has sleeve type bearings, careful attention must be paid to the axial position of the motor shaft in relation to the motor's magnetic centre, when setting the coupling gap.

4)

The compressor external arrangement drawing shows the appropriate connection sizes for lubricating piping connecting to the compressor. It is essential that pipe sizes relate to the connection sizes and that no reduced section pipe fittings are utilised in piping between the manifold and the compressor. Where the connection size is 1" BSP, 1" N.B. piping must be used. It is also essential that oil supply connections on both the male and female rotor sides are piped to the oil manifold as indicated on the external arrangement drawing.

5)

Gas piping must not be allowed to strain the compressor and sufficient pipe supports must be utilised to prevent allowable flange loadings from being exceeded.

6)

Gas suction strainer elements should be rated at 250 microns maximum and consideration should be given to fitting a felt or polypropylene protective sock during initial commissioning runs to prevent contamination from the system pipework. All suction strainer elements must have a free surface area of 2.5 times the suction pipe area and should of the sandwich construction described in Section 8.2.

7)

Oil filter elements should have a minimum rating of 15 microns nominal and should have sufficient surface area to ensure a satisfactory operating life. A duplex filter arrangement is preferred, to allow changeover without stopping the compressor, when low oil/gas differential pressure is experienced.

8)

The oil side of the oil differential above gas discharge pressure must be measured at the oil manifold. Certain operating conditions dictate that an oil/gas differential pressure of 3.0 to 5.0 bar is required and it is essential that alarm and trip settings reflect this higher than normal pressure difference. All WRVT510 compressor selections should be referred to Howden Compressors for confirmation that the oil/gas differential pressure specified is adequate for the operating conditions.

February 2003

9-1


9

SPECIAL INSTRUCTION FOR WRVT510 COMPRESSOR PACKAGING (Cont'd) 9)

Duplex oil pumps, one operating and one standby are preferred. The standby pump should start automatically when the low oil/gas differential pressure alarm set point is reached, thus preventing compressor shutdown.

10)

To allow oil system circulation prior to compressor start up, it is recommended that a solenoid operated valve and a pipe returning to the oil tank/separator vessel be fitted at the end of the oil manifold. This would also allow the oil system to be flushed with the compressor pipes disconnected during precommissioning - See 12. Control is required to ensure that this valve closes and allows the correct oil/gas differential pressure to be established prior to compressor start up.

11)

It is recommended that oil is circulated prior to compressor start up in order to achieve a minimum oil temperature of 30 deg.C. at the compressor bearings. A temperature bypass valve around the oil cooler may be required in order to achieve this.

12)

Cleanliness of the lube oil system is of the utmost importance and must be established prior to any operation of the compressor. During pre-commissioning the pipes between the compressor and the manifold should be disconnected and the system flushed and proved clean before being reconnected to the compressor.

13)

Consideration is required to be given to the relative heights of the compressor and the reservoir oil level to ensure drainage of the compressor on shut down.

14)

To avoid high starting torque caused by the compressor filling with oil prior to start up, consideration should be given to fitting a solenoid valve and a pipe returning to the oil reservoir to the underside connection on the compressor suction casing. As in 10. control is required to ensure that this valve closes prior to compressor start up.

15)

Oil/gas differential pressure must be controlled by a pressure actuated control valve sensing the difference between manifold and discharge pressure. A relief type valve mounted immediately downstream of the oil pump must not be used to control oil/gas pressure difference.

16)

Consideration should be given to fitting pre-alarms and alarms in conjunction with any protective trip switches fitted. This would give adequate warning of any fault condition, thus preventing unnecessary shutdowns and subsequent high transient loads associated with the large masses and inertias in a 510 compressor system.

February 2003

9-2


9

SPECIAL INSTRUCTION FOR WRVT510 COMPRESSOR PACKAGING (Cont'd) 17)

In large refrigeration systems the time taken to reduce the suction pressure to the normal evaporating condition can be considerable. It is essential therefore that the compressor capacity is controlled to a maximum of 50%, until the suction pressure is below the maximum specified in Section 3.1. Any operation of the compressor outside the limits specified in Section 3.1 should be discussed and agreed with Howden Compressors prior to start up.

February 2003

9-3


10

COMPRESSOR SELECTION PROGRAM Compressor selection programs are available upon request to:

Howden Compressors Limited Compressor Business Unit 133 Barfillan Drive Glasgow. G52 1BE, Scotland, UK Tel: 0044 (0)141 882 3346 Fax: 0044 (0)141 882 8648 E-mail: [email protected] Web site: www.howdencompressors.co.uk

or

Howden Compressors LLC 1850B North Gravers Road Plymouth Meeting PA 19462 USA Tel: 001 610 313 9800 Fax: 001 610 313 9215 E-Mail: [email protected] Web site: www.howdencompressors.com

February 2003

10-1

HDWRV Databook

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