CH642721A5 - Method for operating a screw compressor. - Google Patents

Description

642721

2nd

PATENT CLAIMS

1. A method of operating a screw compressor for gases, in which at least two elements move relative to one another with the formation of a sealing gap and an essentially incompressible fluid as a lubricant and sealant seals the gap, characterized in that the operating temperature is in a range from 60 to 80CC is maintained that the main rotor is rotated at a peripheral speed (Uhl) which is between 0.5 to 1.5 times a reference peripheral speed (UoPt) at which the specific compression performance (a) has its minimum that with a sealant flow working is sufficient to maintain the set operating temperature and that a sealant is used, the viscosity, measured at 50 ° C, in the range of 0.5 and 15 cSt.

2. The method according to claim 1, characterized in that at a main rotor reference peripheral speed (Uopt) between 30 and 45 ms-1, the viscosity of the sealing medium is determined at 50 ° C according to the curve according to FIG. 5, deviations of ± 20% of the value determined in this way are permitted.

The delivery rate X l characterizes the utilization of the displacer volume of the compressor by the pumped gas, and is the ratio of the pumped (mf) to the theoretically suckable gas mass (mth):

ÀL =

mf mth

(3)

Further parameters for describing the subject matter of the invention are:

the kinematic viscosity v of the sealing medium to be introduced; Dimension mVs or cSt, and the ratio ß of the volume flow of the sealing medium Vöi to the volume flow of the intake air Van

The starting point for the invention is the requirements that the sealing medium has to meet inside the compressor. They are essentially: lubrication of the mechanically moving parts, cooling of the gas to be compressed and sealing of the work area. As far as "air" is used in the following, this should also be understood to mean another gas or another compressible medium.

In the case of a screw compressor, for example of the SKF102 G type from GHH Sterkrade AG, the sealing medium is introduced near the gas inlet and travels through the compressor while the air is being compressed. The sealing medium is expelled with the air, then separated from it and used again.

In order to check the effects of the properties of the sealing medium used on the function and effectiveness of the compressor, it is necessary to determine the specific compression performance, the grade and the degree of delivery at a certain required volume flow and pressure ratio between air inlet and outlet.

Specific compression performance a is the effective performance, based on the intake volume flow:

a =

Van

(-

kW

mJmin

3T-)

(1)

The following definitions are used:

effective power (coupling power) Pe in kW volume flow, drawn in Van in m3min "'

Accordingly, the parameter a describes the ratio of the clutch power consumed to the volume flow conveyed, based on the intake state.

The quality grade represents the ratio of isentropic compression power Ps to the absorbed clutch power P =:

T] ge =

(2)

ß =

Völ

Van

2o A non-machine characteristic number of the main rotor is a dimensionless parameter y

Y =

Uhl Uopt

25th

introduced. Uhl is the peripheral speed of the main rotor. Uopt as the reference circumferential speed is achieved with every screw compressor when the specific power a is minimal. The reference variable Uopt is different for each compressor and varies with the volume flow Vöi of the sealing medium.

It is known to operate a screw compressor in such a way that, regardless of the drive speed and thus the main rotor peripheral speed Uhl, a seal-3s medium volume flow Vöi with a specific kinematic target viscosity v is introduced into the working space, the cooling capacity being achieved by varying Vöi sufficient to dissipate the heat from the screw compressor at a given compression temperature 40 (eg 80 ° C) of the escaping gases.

According to the specifications of various manufacturers of screw compressors, sealing media are chosen that have a relatively high kinematic viscosity, whereby it was assumed that relatively viscous oil has a cheaper sealing effect. For the operation of screw compressors, minimum viscosities are therefore required to fulfill the sealing tasks. The regulations of various manufacturers are cited as an example:

1. BAUER screw compressor systems ER 20 - ER 75: For air compressors, aging-resistant, water-repellent, non-foaming, anti-corrosion oils with the following data are recommended:

55 base oils viscosity at 50 ° flash point o.T. so pour point

Solvent raffinate approx. 19.5 to 25 cSt above 200 ° C

at least 5 ° below the lowest ambient temperature.

2. IRMER + ELZE screw compressors S 501 - «s SLS: The oil used should have a viscosity of 29 to 33.5 cSt (4-4.5 ° E) at 50 ° C as well as being age and emulsion resistant and only contain normal corrosion, oxidation and anti-foaming agents.

3.Mahl screw compressor BA 231-1:

We recommend particularly age-resistant, water-repellent, non-foaming, anti-corrosion oils with the following data:

Base oil = solvent raffinate Viscosity at 50 ° C = approx. 19.5 to 25 cSt Flash point o.T. = over 200 ° C

Pour point = at least 5 ° C below the lowest ambient temperature.

4. ATLAS COPCO:

For compressor lubrication, the use of a high quality hydraulic oil with anti-rust, anti-foam and anti-oxidation additives with the following viscosity is recommended:

1. Ambient temperature constantly above -10 ° C

3 ° E at 50 ° C (104 saybolt seconds), which corresponds to approximately SAE 10 or SAE10 W.

2. Ambient temperature below 0 ° C

2 ° E at 50 ° C (66 saybolt seconds), which corresponds to approximately SAE 5.

Converted to cSt, the measurement data correspond as follows:

3 ° E or 104 saybolt seconds correspond to approximately 20 cSt ± 5 cSt.

2 ° E or 66 Saybolt seconds correspond to approximately 10CSt ± 2CSt.

However, the last value should only be selected if winter operation is pronounced, so that under normal conditions, such as prevailing in Central Europe, the higher medium viscosity should be selected.

Compressors that are operated with the sealants mentioned generally work satisfactorily. However, it has been shown that when the oil viscosities mentioned are used, the specific output a is relatively high, ie is unfavorable. Furthermore, the relatively highly viscous oils have long separation times and therefore require large-volume separators. The separator, in particular, essentially determines the size and the initial cost of the compressor.

It must therefore be considered disadvantageous for the operation of a screw compressor in accordance with the regulations of the prior art that the oil selection is only made from the point of view of the sealing tasks. This is obviously due to the fact that the degree of delivery Xl was considered the main parameter for given compression designs. However, the degree of delivery is only a mass ratio and therefore cannot give any information about the energy conversion and energy consumption of the compressor. The choice of the degree of delivery as the main parameter therefore obscures the view that other operating parameters allow the optimization of the process to be defined much better.

In order to improve the state of the art, the basic task is to increase the efficiency when operating screw compressors and, taking into account the relative speed of the elements moving relative to one another, to specify selection criteria for the sealing media in order to minimize entropy generation in the compressor. The essential part of entropy generation arises - as plausibility studies show and experiments confirm - when flowing through the sealant

642 721

through the sealing gap. The stated object is achieved by the method according to the invention, which is characterized by the features stated in the characterizing part of patent claim 1.

It has been found that, as a departure from the known regulations for operating a screw compressor, a substantially lower viscosity can in principle be selected for the sealing media used, that the energy losses can thereby be reduced and that the efficiency of the compressor can be improved. In addition, by optimizing the choice of sealant, the size can be reduced by using smaller separators.

A compression ratio% between 5 and 15 is assumed.

Even though the stated values for the viscosity encompass relatively wide limits, the stated values already clearly show that the viscosity is substantially below that which was previously considered desirable for the prior art. It can be demonstrated that by choosing the lower viscosity, a significant reduction in the specific power output can be achieved. It is assumed that a constant compressor end temperature is maintained. The latter can be achieved, for example, by selecting a high volume flow from the sealing medium at a high injection temperature or a low oil flow at a low injection temperature. The selection of the viscosity value can be further optimized, with explanations for this being given in the description of the drawing.

In the following, drawings and examples are attached to explain the invention. The figures show:

1 shows the (simplified) working diagram of a screw compressor system with oil injection cooling and drive by an electric motor;

2 shows an enlarged representation of the conditions in the gap area of two elements which are actually moved relative to one another in a screw compressor;

Fig. 3 is a graph showing the specific performance versus the main rotor peripheral speed of a particular screw compressor for different viscosities of the sealing medium;

Fig. 4 is a graph showing the specific power versus the main rotor peripheral speed;

Fig. 5 is a diagram showing the dependence of the selected viscosity on the main rotor peripheral speed.

In Fig. 1, the scheme of a single-stage screw compressor system with oil injection cooling and driven by an electric motor is shown to understand the function. The center of the system is represented by the compressor 1. A gas, preferably air, is fed to it via an intake filter 3, an intake valve 4 (with a bypass valve 5) under atmospheric pressure. The compressor is driven by an electric motor 2.

A cooling and sealing medium is pressed into the compressor via connection 32 («oil injection»). The sealing medium is supplied via a line 36 from an oil tank 8, which is also a coarse separator. The line 36 is routed through the usual operating stations, namely via check valve 24, dirt trap 25, oil cooler 26, oil filter 28 and throttle valve 30 and check valve 31.

A line branches off from line 36 to connections 35 (“gear lubrication”) or 33 (“relief piston”). The oil tank 8 also has an oil filler neck 21, an oil drain 23 and a level indicator 22.

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4th

The compressed gas is fed to the oil tank and separator 8 via an outlet line 37 (with check valve 7) and fed to a consumer via a line 38 with fine separator 9 and check valve 13.

It is known to use mineral or silicone oils as coolants, sealants and lubricants. The oil flow brought into the compressor room has three main effects:

a) lubrication;

b) workspace seal, i.e. Reduction of the air backflow through an axial and radial seal through which an oil-air mixture flows;

c) heat transfer.

It is based on a model that says that apart from the “blowhole”, all edges that seal the work space push an oil surge in front of them during a certain time interval due to their sealing movement in the fixed work space. If this surge of oil persists during an entire work cycle, the sealing gap flows through oil in the direction of the pressure drop. The resulting pressure losses (entropy generation through dissipation) are essentially determined by the oil viscosity of an as yet unknown flow law, the further parameters of which, in addition to the gap geometry, are the gap pressure and the working length.

With increasing working length (decreasing speed) the hydraulic losses in the machine decrease, and at the same time the specific compression performance decreases with decreasing oil viscosity as long as the sealing gaps are filled with backflowing oil.

If this is no longer the case as the working length increases, i.e. If the sealing gaps are only filled during certain working time intervals, i.e. if there are times when oil-air mixtures or only air flow through the gaps, energy is lost. The returning air has been compressed beforehand, but its energy content is lost again, so that the specific output increases.

To explain the energy loss sources, the conditions at a sealing gap are shown in an enlarged view in FIG. 2. An edge 40 of the compressor screw rotating in a housing moves in the direction of arrow 41 with a rotating movement along the inside 42 of the housing. A gap 43 is formed between the elements 40 and 42. The gap is filled with a part of the sealing medium 44, which is pushed as a gush in front of the gap and is also located in the gap itself and behind it. The losses resulting from rolling and sliding movements due to the toughness of the sealing medium are referred to as dissipation losses. At the same time, as indicated by the dashed arrow 44 ', there is a leakage loss of the highly compressed gases which pass through the sealing medium and cause a pressure loss. In order to arrive at a comparison and selection criteria for the sealing medium to be selected, tests were carried out, the results of which are reflected in FIGS. 3-5 and the examples.

example 1

(see Fig. 3)

A screw compressor type SKF102 G with a pair of rotors (manuf .: GHH Sterkrade AG) is operated with the following values:

Pressure ratio% = 8

Compressor end temperature t2 = 80 ° C

Front gap 10-2mm

Series play = 6.1 • 10-2 mm

Oil flow Vöi = 43 1 min-1.

Specific power measurements are made as a function of the peripheral speed Uhl of the main rotor at the following speeds:

15, 20, 25, 30, 35 and 45 ms "1.

The resulting curves are shown in FIG. 3. They have been measured with hydraulic oils of the following viscosities:

H-L68 68 cSt at 50 ° C

H-L25 25 cSt at 50 ° C H-Lll llcSt at 50 ° C

The low viscosity influence on the specific performance at low peripheral speeds is remarkable. Long working hours imply long dwell times of the suctioned load in the closed work space, which can only be insufficiently sealed due to Vöi = 43 1 min-1. Since there is no oversupply of work space sealing oil, the change in oil viscosity has a particularly significant impact. The result is an increase in the degree of delivery (not shown). The minima of the specific capacities shift slightly when the viscosity decreases towards higher peripheral speeds.

From this arbitrarily selected example it can already be seen that there is an improvement in performance and efficiency through the use of low-viscosity oils.

Example 2

(see FIGS. 4 and 5)

In order to eliminate the parameter Vöi (oil flow), a variable Uhl is introduced. It is assumed that

= f (UHL) is>

i.e. a function of the main rotor peripheral speed, a minimum of the quotient a to which a specific Uhl is assigned can be found for each oil flow Vöi that is suitable for maintaining a specific compressor end temperature t2. By definition, this variable is defined as Uopt, i.e. as the optimal main rotor peripheral speed or as the reference peripheral speed.

By varying Vöi, family of curves y can be determined as a function of UoPt. 4 shows five curves which were obtained with different viscosities of the sealing medium:

Curve 68: HL 68 oil with 68 cSt at 50 ° C Curve 25: HL 25 oil with 25 cSt at 50 ° C Curve 14: HL 14 oil with 14 cSt at 50 ° C Curve 10: Shell Tellus 117 (10 cSt at 50 ° C )

Curve 4: Shell Oil T17 (4 cSt at 50 ° C)

It turns out that when shifting to high peripheral speeds, with appropriate control of the sealant flow with the only criterion to keep the final temperature at a certain value between 60 ° and 90 ° C, a significant improvement in the specific performance compared to the prior art can be achieved is. According to 5 in the area of interest 30 ms-1 <Uopt <45 ms-1, the values are punctiform, a parabola or hyperbola results in the first approximation, which can be defined as follows:

s

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5

642721

v [cSt] = a IPopi [m2s ~ 2] + b Uopi [ms_1] + c where a = + 0.0564,

b = - 5.11 and c = + 119.8.

Using the curve according to FIG. 5, operating modes that differ from the regulations and operating modes of the prior art can be determined, which ensure an improved efficiency of the screw compressor. The viscosity values determined from the curve can be used upwards or downwards deviating by approx. 10-20% - which is necessary for the reason alone, since only oils of certain viscosities are available on the market. The main rotor peripheral speed plays an important role. At medium and high circumferential speeds, there is a proportional decrease in specific power with decreasing viscosity. Depending on the selected conditions, however, a 4 cSt oil can bring worse conditions than a 14 cSt oil, io if the optimal number of revolutions is significantly undercut. An oil that can be used as universally as possible should therefore have a viscosity between 10 and 14 cSt.

B

4 sheets of drawings